Toroidal continuously variable transmission

ABSTRACT

The object of the present invention is to provide compact toroidal type continuously variable transmission for a front engine front wheel drive vehicle, which employs geared neutral starting system. The toroidal type continuously variable transmission according to the present invention comprises an input shaft  11  whose one end is coupled with an engine, a primary shaft  12  into which said input shaft is fittingly inserted with play, a secondary shaft  13  which is disposed parallel with said shafts and whose engine side end is coupled with a differential gear unit  5 , toroidal type continuously variable transmission mechanisms  20, 30  which are disposed on said primary shaft  12 , a planetary gear mechanism  50  which is disposed on said secondary shaft  13 , a low mode gear train  80  for coupling the input shaft  11  with a carrier  51  of the planetary gear mechanism  50 , and a high mode gear train  90  for coupling a sun gear  52  with an output disk  34  of the continuously variable transmission mechanisms  20, 30,  wherein said low mode gear train  80  is disposed on an opposite side end, with respect to the engine, of the input shaft  11.

BACKGROUND OF THE INVENTION

The present invention relates to a toroidal type continuously variabletransmission, in particular, to a structure of continuously variabletransmission for a front engine front wheel drive vehicle employing ageared neutral starting system.

DESCRIPTION OF THE PRIOR ART

A toroidal type continuously variable transmission, which continuouslychanges a speed ratio of power transmission between an input disk and anoutput disk by interposing a roller for making the power transmissionbetween both disks and by rotating this rotor with a variableinclination angle so as for contact points thereof with each of saidboth disks to be changed in a radial direction respectively, is nowbeing introduced into practical use as a continuously variabletransmission for automobile, and, as shown, for example, in the officialgazette of Japanese Patent Un-examined Publication No. 6-101754, ageared neutral starting system has been proposed to use in this typecontinuously variable transmission.

In this system, the continuously variable transmission mechanism havingthe structure described above is disposed on a first shaft coupled withan engine, and a planetary gear mechanism comprising three rotaryelements, that is, a sun gear, an internal gear, and a pinion carrierfor supporting a planetary pinion engaging with these both gears, isalso disposed on a second shaft which is parallel with the first shaft,wherein a revolution of the engine is transmitted to the first rotaryelement directly, and to the second rotary element through thecontinuously variable transmission mechanism mentioned above and thethird rotary element is made to be a power output element.

A neutral condition is realized by controlling the speed ratio of thecontinuously variable transmission mechanism so that a ratio of inputrotation speeds between the first and the second rotary elements of theplanetary gear mechanism is controlled so as to stop the third rotaryelement, and, starting form this condition, the third rotary element,which is the power output element, is rotated in the forward or thereverse direction by increasing or decreasing the speed ratio of thecontinuously variable transmission mechanism.

This system allows the vehicle to be started without using a clutch, atorque converter and the like, and this improves a response in startingand an efficiency of power transmission.

When above mentioned geared neutral system is employed, two powertransmission paths are necessary, one of which transmits the power fromthe first shaft side to the planetary gear mechanism on the second shaftwithout interposing the continuously variable transmission mechanismtherebetween, and the other of which transmit the power from the firstshaft side to the planetary gear mechanism through the continuouslyvariable transmission mechanism, and therefore, in the continuouslyvariable transmission disclosed in the above official gazette, thelatter power transmission path is provided in the middle of a tandemtype transmission gear mechanism and the former power transmission pathwhich builds up a gear train for a low mode reduction gear system isprovided in the engine side of the continuously variable transmissionmechanism.

In this type of structure, however, especially in the transmission forthe front engine front wheel drive vehicle, there occurs problems asbelow.

That is, in the transmission for the front engine front wheel drivevehicle, as shown in FIG. 3 of the above gazette, a differential gearunit is generally connected to an engine side end of the second shaft (,which is an output shaft of a transmission unit) on which the planetarygear mechanism is disposed, so that, in this case, the gear train fordirectly transmitting the power from the first shaft side to theplanetary gear mechanism on the second shaft must be disposed near bythe differential gear unit having a large diameter. Accordingly, inorder to avoid an interference between the gear train and thedifferential gear unit, they must be placed with some distancetherebetween with respect to the shaft direction, which increases thesize of the transmission along the shaft direction. Especially in thefront engine front wheel drive vehicle which has a shaft line of theengine and the transmission disposed laterally with respect to thevehicle center line, the increased size in this shaft line directionmakes a mounting operation to a vehicle body as well as a layout designof the engine and the transmission difficult.

Accordingly, the object of the present invention is to increase anallowance for the layout and the mounting operation to the vehicle bodyof the toroidal type continuously variable transmission of the frontengine front wheel drive vehicle employing the geared neutral startingsystem by controlling the increase of the size in the shaft direction.

SUMMARY OF THE INVENTION

In order to solve the problem mentioned above, the present invention ischaracterized by the constitution described below.

In an invention claimed in claim 1 of the present invention (hereafterreferred to as the first invention), a first shaft engaged with anengine on one end thereof and a second shaft disposed parallel with saidfirst shaft and engaged, on an engine side end thereof, with adifferential gear unit for driving a left and a right driven wheel areprovided, and on said first shaft is disposed a toroidal typecontinuously variable transmission mechanism comprising an input diskcoupled with said first shaft, an output disk disposed in the engineside of said input disk and rotatably supported by the first shaft, aroller interposed between these disks for transmitting a power betweenboth disks, and a contact point control member for changing a speedratio between both disks by inclinably and rotatably supporting saidroller and thereby changing the contact point between said roller andeach of the input and the output disks, and on said second shaft isdisposed a planetary gear mechanism comprising three rotary elements,that is, a sun gear, an internal gear and a pinion carrier, wherein,among these rotary elements, the first element is coupled with theoutput disk of the continuously variable transmission mechanism so as torotate together with it and the second element is coupled with thesecond shaft.

Above described system is characterized in that a gear train comprisinga first gear disposed on the opposite side with respect to the engine ofthe continuously variable transmission mechanism mounted on the firstshaft so as to rotate together with said first shaft, a second gearrotatably supported on the opposite side with respect to the engine ofthe planetary gear mechanism mounted on the second shaft, and an idlegear which is engaged with these gears to transmit a power between bothgears is further provided, and furthermore a first clutch mechanism forengaging or disengaging the second gear of this gear train with thethird element of the planetary gear mechanism, a second clutch mechanismfor engaging or disengaging the output disk of the continuously variabletransmission mechanism with the second shaft, and a control means forcontrolling operations of said first and said second shaft, and acontrol means for controlling operations of said first and said secondclutch mechanism and the contact point control member are provided.

Preferably, a vehicle speed detecting means is further provided, andsaid control means is characterized in that it controls said firstclutch so as to engage said second gear with said third element and atthe same time controls said second clutch so as to intercept said powertransmission path while the vehicle speed being lower than apredetermined vehicle speed, and controls the first clutch so as tointercept the engagement between the second gear and the third elementand at the same time controls the second clutch so as to engage thepower transmission path while the vehicle speed being higher than saidpredetermined vehicle speed.

In addition, an engine load detector is preferably provided so that saidpredetermined speed is increased as the engine load increases.

In another aspect of the invention, adding to a first continuouslyvariable transmission mechanism comprising the input disk coupled withthe first shaft, the output disk disposed in the engine side of saidinput disk and rotatably supported by the first shaft, the rollerinterposed between these disks, and the contact point control member forchanging the contact point between said roller and the respective disks,the toroidal type continuously variable transmission mechanism furthercomprises a second continuously variable transmission mechanismcomprising a second output disk disposed on the engine side of theoutput disk of the first continuously variable transmission mechanismand rotatably supported by the first shaft, a second input disk disposedon the engine side of said output disk and coupled with the first shaft,a second roller interposed between these disks, and a second contactpoint control member for changing the contact point between said rollerand the respective disks.

The toroidal type continuously variable transmission mechanism describedabove is characterized in that the output disk of the first continuouslyvariable transmission mechanism and the output disk of the secondcontinuously variable transmission mechanism are integrated into oneunit and a gear is formed on an outer surface of said integrated outputdisk unit for interlocking and rotating said integrated output disk andthe first element of the planetary gear mechanism with each other.

In further aspect of the invention, same as the second inventiondescribed above, adding to said first continuously variable transmissionmechanism comprising the input disk coupled with the first shaft, theoutput disk disposed in the engine side of said input disk and rotatablysupported by the first shaft, the roller interposed between these disks,and the contact point control member for changing the contact pointbetween said roller and the respective disks, the toroidal typecontinuously variable transmission mechanism further comprises thesecond continuously variable transmission mechanism comprising thesecond output disk disposed on the engine side of the output disk of thefirst continuously variable transmission mechanism and rotatablysupported by the first shaft together with said output disk integrally,the second input disk disposed on the engine side of said output diskand coupled with the first shaft, the second roller interposed betweenthese disks, and the second contact point control member for changingthe contact point between said roller and the respective disks.

The toroidal type continuously variable transmission mechanism describedabove is characterized in that said first shaft is inserted into athrough-hole formed in a third shaft to dispose each of the input sidedisks and the output side disks of said first and said secondcontinuously variable transmission mechanisms on said third shaft,wherein one end of said third shaft is supported by a transmission casethrough a bearing and onto the other end thereof is fitted into a firstgear of a gear train, said first gear being supported by thetransmission case through a bearing, and further a spring member isinterposed in the fitting portion between said third shaft and the firstgear to absorb relative displacement therebetween in the shaft linedirection.

In still further aspect of the invention, same as the third inventiondescribed above, adding to said first continuously variable transmissionmechanism comprising the input disk coupled with the first shaft, theoutput disk disposed in the engine side of said input disk and rotatablysupported by the first shaft, the roller interposed between these disks,and the contact point control member for changing the contact pointbetween said roller and the respective disks, the toroidal typecontinuously variable transmission mechanism further comprises thesecond continuously variable transmission mechanism comprising thesecond output disk disposed on the engine side of the output disk of thefirst continuously variable transmission mechanism and rotatablysupported by the first shaft, the second input disk disposed on theengine side of said output disk and coupled with the first shaft, thesecond roller interposed between these disks, and the second contactpoint control member for changing the contact point between said rollerand the respective disks.

The toroidal type continuously variable transmission mechanism describedabove is characterized in that said first shaft is inserted into athrough-hole formed in a third shaft to rotatably support both outputdisks of the first and the second continuously variable transmissionmechanisms integrally with each other on the middle of said third shaft,and, on an opposite side, with respect to the engine, and on the engineside of said both disks, input disks of the first and the secondcontinuously variable transmission mechanisms are disposed on and areengaged with said third shaft respectively, wherein a loading mechanismfor pressing the roller by and between the input and the output disks inthe first and the second continuously variable transmission mechanismsis provided between the input disk of the first continuously variabletransmission mechanism and the first gear of the gear train disposed onthe opposite side thereof with respect to the engine.

In yet further aspect of the invention, the toroidal type continuouslyvariable transmission mechanism is characterized in that the loadingmechanism comprises a pair of disks whose surfaces facing with eachother are formed into cam surfaces having circumferential concave andconvex and a roller which is interposed between both disks to generateaxial force between them by the relative rotation therebetween, and apin member is interposed between the first gear of the gear train andthe disk located in said first gear side to integrally rotate them,wherein said pin member is disposed in a portion where a thickness ofthe disk located in said first gear side is rather thicker due to theconcave and convex figures thereof.

Further, in the first invention described above, two oil channels forsupplying the first clutch mechanism and the second clutch mechanismwith a coupling fluid respectively are provided in the second shaft,wherein said both oil channels may be led from a side portion where ahydraulic pressure source is provided.

According to the structure described above, following operations may beachieved.

First, when the first clutch mechanism is engaged, that is, the secondgear of the gear train is engaged with the third element of theplanetary gear mechanism, and the second clutch mechanism is disengaged,that is, the output disk of the continuously variable transmissionmechanism is disengaged from the second shaft, the revolution inputtedfrom the engine into the first shaft is inputted from said first shaftthrough the gear train and the first clutch mechanism into the thirdelement of the planetary gear mechanism disposed on the second shaft andis also transmitted from the input disk through the roller to the outputdisk in the continuously variable transmission mechanism on the firstshaft and then is inputted form said output disk into the first elementof said planetary gear mechanism.

At that time, if the speed ratio of the continuously variabletransmission mechanism is appropriately controlled by the control meansthrough the contact point control member so that the rotation speedratio between the first and the third elements of said planetary gearmechanism is set so as for the rotation speed of the second elementthereof to be zero, the rotation of the second shaft, which is theoutput shaft of this transmission, can be stopped while the enginerevolution being inputted and the first clutch mechanism being engage,that is, the geared neutral condition can be achieved.

When, staring from this condition, the speed ratio of this continuouslyvariable transmission is changed so that the rotation speed of the firstelement of the planetary gear mechanism is increased or decreased, thesecond shaft is rotated in the forward or the backward running directionunder a low mode condition, in which the speed ratio as a wholetransmission is large, that is, the vehicle will start.

When the first clutch mechanism is disengaged, that is, the second gearof the gear train is disengaged from the third element of the planetarygear mechanism, and the second clutch mechanism is engaged, that is, theoutput disk of the continuously variable transmission mechanism isengaged with the second shaft, the revolution inputted from the engineinto the first shaft is transmitted from the continuously variabletransmission mechanism through only the second clutch mechanism into thesecond shaft. At that time, since the planetary gear mechanism does notchange a speed ratio, the speed ratio as a whole transmissioncorresponds to that of the continuously variable transmission mechanism,and this means that the speed ratio is controlled under so-called highmode, where the speed ratio is small, by the continuously variabletransmission mechanism without any steps but continuously.

Since the gear train which transmits the rotating motion from the firstshaft to the planetary gear mechanism under the geared neutral conditionor the low mode condition is disposed on the opposite side, with respectto the engine, of the continuously variable transmission mechanism onthe first shaft and the planetary gear mechanism on the second shaft,this gear train is prevented from interfering with the differential gearunit which is engaged with the second shaft on its end of engine side,and this allows the length of the transmission unit along the shaft linedirection to be shortened.

Further, in the case where the first and the second continuouslyvariable transmission mechanisms are provided as a toroidal typecontinuously variable transmission mechanism comprising a pair of inputand output disks, a roller interposed between both disks and the like,since the output disks of the first and the second continuously variabletransmission mechanisms are integrated into one unit and the gear isformed on the outer surface thereof for engaging and rotating saidintegrated output disk together with the first element of the planetarygear mechanism, the length along the shaft line direction is madeshorter and the gear may be supported more stably, which prevents thebacklash of the gear along the shaft line direction, comparing with thecase where two output disks are provided independently and said gear isdisposed between said both disks.

Furthermore, same as the case described above, in the case where thefirst and the second continuously variable transmission mechanisms areprovided as a toroidal type continuously variable transmissionmechanisms and the first shaft is inserted into the through-hole formedin the third shaft to dispose each of the input side and the output sidedisks of said first and said second continuously variable transmissionmechanisms on said third shaft, since the one end of said third shaft issupported by the transmission case through the bearing and onto theother end thereof is fitted into the first gear of the gear train, saidfirst gear being supported by the transmission case through a bearing,and further a spring member is interposed in the fitting portion betweensaid third shaft and the first gear to absorb relative displacementtherebetween in the shaft line direction, the expansion and contractionof the third shaft can be absorbed by said spring member even if thethird shaft is expanded or contracted due to the thermal expansion andthe like.

Therefore, an axial force applied to the bearings one of which supportsone end of the third shaft and the other of which supports the other endof the third shaft through the first gear is maintained properly and theaxial backlash of the first gear is also controlled, so that the firstgear can be maintained in good condition. On the other hand, since, inthe case where the first and the second continuously variabletransmission mechanisms are provided as a toroidal type continuouslyvariable transmission mechanism and the first shaft is inserted into thethrough-hole formed in the third shaft on the middle of which arerotatably supported the output disks of the first and the secondcontinuously variable transmission mechanisms and, on the opposite sidewith respect to the engine and on the engine side thereof, the input andthe output disks of the first and the second continuously variabletransmission mechanisms are disposed and are engaged with the thirdshaft respectively, since the loading mechanism for respectivelyapplying pressure onto the rollers by and between the input and theoutput disks in the first and the second continuously variabletransmission mechanisms is disposed between the input disk of the firstcontinuously variable transmission mechanism and the first gear of thegear train disposed on the opposite side thereof with respect to theengine, a torque flow from the engine under the geared neutral or thelow mode conditions where the first clutch mechanism is engaged and thesecond clutch mechanism is disengaged can be appropriately carried out.

Under this condition, the torque from the engine is inputted into thefirst shaft, then is transmitted from the opposite side end with respectto the engine of the first shaft through the tear train to the secondshaft side, and then is inputted through the first clutch mechanism intothe third element of the planetary gear mechanism. At that time, in thisplanetary gear mechanism, the torque is outputted from the secondelement through the second shaft to the differential gear unit side andsimultaneously a reaction force against the torque input into the thirdelement is applied to the first element, and then this reaction force iscirculated back to the output disks of the first and the secondcontinuously variable transmission mechanisms and thereby the so-calledcirculating torque is generated.

As for this circulating torque, some part thereof transmitted to theinput disk of the first continuously variable transmission mechanism istransmitted through the loading mechanism to the first gear of the geartrain, and the other part thereof transmitted to the input disk of thesecond continuously variable transmission mechanism is transmittedthrough the third shaft from the loading mechanism, in the same manner,to the first gear of the gear train respectively. Accordingly, eithercirculating torque does not pass through the first shaft, so that thefirst shaft is required to make only the engine torque pass through.Further, since, in the case where the loading mechanism comprises a pairof disks whose surfaces facing with each other are formed into the camsurfaces having circumferential concave and convex and a roller which isinterposed between both disks, and the disk of the first gear side ofthe loading mechanism and the first gear are jointed by the pin member,said pin member is disposed in the portion of the disk of the first gearside where the thickness of the disk is rather thicker due to theconcave and convex figures thereof, the first gear can be coupled withthe disk without increasing the thickness of the disk, that is, thelength along the shaft line direction, as a whole, without decreasingthe strength of the disk.

Furthermore, since, when two oil channels for supplying the first andthe second clutch mechanisms with the coupling fluid respectively areprovided in the second shaft on which both clutch mechanisms aredisposed, both oil channels are led from the side portion where thehydraulic pressure source is disposed, the length of the oil channels toboth clutch mechanisms is made short and thereby the hydraulic pressuremay be supplied to these clutch mechanism quickly, which makes improvedresponse of engagement and disengagement control.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic skeleton illustrating a mechanical structure ofthe toroidal type continuously variable transmission of the preferredembodiment according to the present invention;

FIG. 2 is a development illustrating the concrete structure of the mainpart of said transmission;

FIG. 3 is a cross sectional view taken on line A—A of FIG. 2;

FIG. 4 is a cross sectional view illustrating an assembling aspect ofthe gear constructing the high mode gear train;

FIG. 5 is a partially sectional view illustrating an assembling relationof the loading cam, the gear constructing the low mode gear train, andthe input disk;

FIG. 6 is an enlarged cross sectional view illustrating a structure onthe input shaft;

FIG. 7 is an enlarged cross sectional view illustrating a structure onthe secondary shaft;

FIG. 8 is a schematic diagram for explaining the problem due to thecirculating torque;

FIG. 9 is a schematic diagram for explaining the flow of the circulatingtorque in the transmission of the preferred embodiment according to thepresent invention;

FIG. 10 is a circuit diagram for the hydraulic pressure control of saidtransmission;

FIG. 11 is a partially sectional view taken in the direction of arrow Bof FIG. 3 illustrating the three-layers valve for generating hydraulicpressure for shift control;

FIG. 12 is a partially sectional view taken in the direction of arrow Cof FIG. 3 illustrating the cam mechanism;

FIG. 13 is a cross sectional view illustrating the lower structure ofthe transmission case;

FIG. 14 is a control system diagram of the transmission of the preferredembodiment according to the present invention;

FIG. 15 is an explanatory drawing for the traction force as aprecondition of the shift control;

FIG. 16 is a characteristic drawing illustrating a relation between thenumber of the pulse of the step motor and the toroidal speed ratio;

FIG. 17 is a characteristic drawing illustrating a relation between thenumber of the pulse of the step motor and the final speed ratio;

FIG. 18 is a characteristic diagram used in the shift control;

FIG. 19 is an explanatory drawing for the problem in the shift controlby the three-layers valve;

FIG. 20 shows the main flow chart implemented by the control unit;

FIG. 21 is an explanatory drawing illustrating the feature of the linepressure control implemented by said control unit;

FIG. 22 is a flow chart of said line pressure control;

FIG. 23 is a characteristic diagram of said line pressure control;

FIG. 24 is another characteristic diagram of said line pressure control;

FIG. 25 is a flow chart of the engage control implemented by saidcontrol unit;

FIG. 26 is a characteristic diagram of said engage control;

FIG. 27 is another characteristic diagram of said engage control;

FIG. 28 is a flow chart of the direct control implemented by saidcontrol unit;

FIG. 29 is a characteristic diagram of said direct control;

FIG. 30 is another characteristic diagram of said direct control;

FIG. 31 is a time chart of said direct and engage controls;

FIG. 32 is a flow chart of the second direct control including theinclination control;

FIG. 33 is a characteristic table of said second direct control;

FIG. 34 is a characteristic diagram of said second direct control;

FIG. 35 is a time chart of said second direct control;

FIG. 36 is a flow chart of the switching control implemented by saidcontrol unit;

FIG. 37 is another flow chart of the switching control;

FIG. 38 is a flow chart of the shift control in reverse implemented bysaid control unit;

FIG. 39 is a shift-characteristic diagram of said shift control inreverse;

FIG. 40 is a flow chart of the mode-switching control implemented bysaid control unit; and

FIG. 41 is a characteristic diagram of said mode-switching control.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

There will now be described a mechanical structure, a structure of thehydraulic pressure control circuit, and a concrete operation of theshift control of the continuously variable transmission according to thepresent invention.

Mechanical Structure

FIG. 1 is a schematic skeleton illustrating a mechanical structure ofthe toroidal type continuously variable transmission of the preferredembodiment according to the present invention, in which the transmission10 comprises an input shaft (the first shaft) 11 coupled with an outputshaft 2 of an engine 1 through a torsional damper 3, a primary shaft(the third shaft) 12 having a through-hole inside thereof into whichsaid shaft 11 is inserted, and a secondary shaft (the second shaft) 13disposed parallel with said shafts 11 and 12, wherein all of said shafts11 to 13 are disposed so as to extend in the lateral direction of thevehicle.

Further, in this transmission 10, a toroidal type first continuouslyvariable transmission mechanisms 20, a toroidal type second continuouslyvariable transmission mechanisms 30, and a loading cam 40 are disposedon a shaft line of said input shaft 11 and said primary shaft 12, and aplanetary gear mechanism 50, a low mode clutch (the first clutchmechanism) 60, and a high mode clutch (the second clutch mechanism) 70are disposed on a shaft line of said secondary shaft 13. A low mode geartrain 80 and a high mode gear train 90 are interposed between the shaftline of the input shaft 11 and the primary shaft 12 and the shaft lineof the secondary shaft 13.

The first and the second continuously variable transmission mechanisms20, 30 have similar structure with each other, in which each of them hasan input disk 21, 31 and an output disk 22, 32, each having facingsurface formed into toroidal surface respectively, and two pieces ofrollers 23, 33 are respectively interposed between said respective bothdisks 21, 22 and 31, 32 to transmit a driving force between themrespectively.

In the first continuously variable transmission mechanism 20 disposed inthe distal side from the engine, the input disk 21 is disposed in thedistal side from the engine and the output disk 22 is disposed in theproximal side from the engine, and in the second continuously variabletransmission mechanism 30 disposed in the proximal side from the engine,the input disk 31 is disposed in the proximal side from the engine andthe output disk 32 is disposed in the distal side from the engine, andfurther, the input disks 21, 31 of respective continuously variabletransmission mechanisms 20, 30 are coupled with respective ends of theprimary shaft 12, and the output disks 22, 32 thereof are formed intoone unit and are rotatably supported on the middle of said primary shaft12.

A first gear 81 included in the low mode gear train 80 is coupled with adistal end, from the engine, of the input shaft 11, and a loading cam 40is interposed between said first gear 81 and the input disk 21 of thefirst continuously variable transmission mechanism 20, and further, afirst gear 91 included in the high mode gear train 90 is provided on anouter surface of the integrated output disks 22, 32 (hereafter referredto as a “integrated output disk 34”) of the first and the secondcontinuously variable transmission mechanisms 20, 30.

On the other hand, a second gear 82 included in the low mode gear train80 is rotatably supported on a distal end from the engine of thesecondary shaft 13 and is engaged with the first gear 81 through an idlegear 83, and further, the planetary gear mechanism 50 is disposed on themiddle of the secondary shaft 13. Between a pinion carrier (the thirdrotary element) 51 of the planetary gear mechanism 50 and the secondgear 82 of the low mode gear train 80 is disposed a low mode clutch 60for engaging or disengaging them with each other.

A second gear 92 engaged with the first gear 91 of the high mode geartrain 90, which is provided on the outer surface of the integratedoutput disk 34 of the first and the second continuously variabletransmission mechanisms 20, 30, is rotatably supported on the engineside of the planetary gear mechanism 50, and further, said second gear92 is coupled with a sun gear (the first rotary element) 52 of theplanetary gear mechanism 50 and an internal gear (the second rotaryelement) 53 of the planetary gear mechanism 50 is connected to thesecondary shaft 13, and a high mode clutch 70 for engaging ordisengaging the second gear 92 of the high mode gear train 90 with thesecondary shaft 13 is disposed on the engine side of the planetary gearmechanism 50.

A differential gear unit 5 is coupled with the engine side end of thesecondary shaft 13 through an output gear train 4 comprising a first anda second gears 4 a, 4 b and an idle gear 4 c, and a driving force istransmitted to the left and the driven wheels (not shown) through driveshaft 6 a, 6 b extending from the differential gear unit 5 to the leftand the right sides.

Then will now be described a detailed description of each component ofthe transmission 10 with reference to the drawings, FIG. 2 and thefollowings. As for the first and the second continuously variabletransmission mechanisms 20, 30 these first and the second continuouslyvariable transmission mechanisms 20, 30 have similar structure with eachother, in which, as described above, each of them has the input disk 21,31 and the output disk 22, 32 (integrated output disk 34), each havingfacing surface formed into toroidal surface respectively, and two piecesof rollers 23, 33 are respectively interposed between said respectiveinput and output disks 21, 22 and 31, 32 to transmit a driving forcebetween them respectively.

To make a detailed description, for example, of the first continuouslyvariable transmission mechanism 20 with reference to FIG. 3, a pair ofrollers 23, 23 is supported by trunnions 25, 25 through shafts 24, 24extending approximately in the radial direction of the input and theoutput disks 21, 22, and respective rollers are disposed on the toroidalsurfaces of the input and the output disks 21, 22, which are facing witheach other, at opposite side thereof by 180 degree with approximatelyhorizontal attitude and parallel with each other, and are respectivelybrought into contact with the toroidal surfaces of said both disks 21,22 at two portions located in the opposite side with each other by 180degree.

Said trunnions 25, 25 are supported by and between left and rightsupport members 26, 26 which are attached to a transmission case 100,and are allowed to rotate around the horizontal axial center line X, Xwhich is of a tangential direction of both disks 21, 22 and normal tothe shaft 24, 24 of the roller 23, 23 and are also allowed to linearlymove reciprocating along said axial center line X, X direction. A rod27, 27 extending toward one side along said axial center line X, X iscoupled with the trunnion 25, 25, and a shift control unit 110 whichincludes the roller 23, 23 through the rod 27, 27 and the trunnion 25,25 is attached on the side of the transmission case 100.

The shift control unit 110 comprises a hydraulic control section 111 anda trunnion control section 112, wherein a piston for increasing speed1131 and that for decreasing speed 1141, which are attached to the rod27 of a first trunnion 251 located upper side, and a piston forincreasing speed 1132 and that for decreasing speed 1142, which areattached to the rod 27 of a second trunnion 252 located lower side, aredisposed in said trunnion control section 112, and a hydraulic pressurechamber for increasing speed 1151 and that for decreasing speed 1161 areprovided on the facing sides of the upper pistons 1131 and 1141respectively and also a hydraulic pressure chamber for increasing speed1152 and that for decreasing speed 1162 are provided on the facing sidesof the lower pistons 1132 and 1142 respectively.

As for the first trunnion 251 located upper side, the hydraulic pressurechamber for increasing speed 1151 and that for decreasing speed 1161 areprovided on the roller 23 side and on the opposite side thereofrespectively, and as to the second trunnion 252 located lower side, thehydraulic pressure chamber for decreasing speed 1161 and that forincreasing speed 1152 are provided on the roller 23 side and on theopposite side thereof respectively.

Hydraulic pressure for increasing speed PH generated by the hydraulicpressure control section 111 is supplied through an oil channel 117, 118to the hydraulic pressure chamber for increasing speed 1151 of the firsttrunnion 251 located upper side and the hydraulic pressure chamber forincreasing speed 1152 of the second trunnion 252 located lower side, andalso hydraulic pressure for decreasing speed PL generated by thehydraulic pressure control section 111 is supplied through a not-shownoil channel to the hydraulic pressure chamber for decreasing speed 1161of the first trunnion 251 located in the upper side and the hydraulicpressure chamber for decreasing speed 1162 of the second trunnion 252located in the lower side.

Then, the relation between a supply control of the hydraulic pressurefor increasing speed PH as well as that for decreasing speed PL and ashift operation of the continuously variable transmission mechanism 20will be briefly described by taking the first continuously variabletransmission mechanism 20 as an example.

First, when the hydraulic pressure for increasing speed PH supplied tothe hydraulic pressure chambers for increasing speed 1151, 1152 of thefirst and the second trunnions 251, 252 becomes relatively higher thanthe predetermined neutral condition comparing with the hydraulicpressure for decreasing speed PL supplied to the hydraulic pressurechambers for decreasing speed 1161, 1162 of the first and the secondtrunnions 251, 252 due to the operation of the hydraulic pressurecontrol section 111 shown in FIG. 3, the first trunnion 251 in the upperside horizontally moves to the right and the second trunnion 252 in thelower side horizontally moves to the left on the drawing.

At that time, assuming that the shown output disk 22 is rotating in thex-direction, a downward force is applied to the upper first roller 231,due to the movement to the right, by the output disk 22 and an upwardforce is applied thereto by the input disk 21 which is located this sideof the paper and is rotating in the opposite of x-direction. To thelower second roller 232, due to the movement to the left, an upwardforce is applied by the output disk 22 and a downward force is appliedby the input disk 21. As a result, both of the upper and the lowerrollers 231, 232 incline so that the contact points with the input disk21 move to the outer side in the radial direction and that with theoutput disk 22 moves to the inner side in the radial direction, andconsequently the speed ratio of the continuously variable transmissionmechanism 20 becomes smaller (increase of speed). On the contrary, whenthe hydraulic pressure for decreasing speed PL supplied to the hydraulicpressure chambers for decreasing speed 1161, 1162 of the first and thesecond trunnions 251, 252 becomes relatively higher than thepredetermined neutral condition comparing with the hydraulic pressurefor increasing speed PH supplied to the hydraulic pressure chambers forincreasing speed 1151, 1152 of the first and the second trunnions 251,251, the first trunnion 251 in the upper side horizontally moves to theleft and the second trunnion 252 in the lower side horizontally moves tothe right on the drawing.

At that time, to the upper first roller 231, an upward force is appliedby the output disk 22 and a downward force is applied by the input disk21, and, to the lower second roller 232, a downward force is applied bythe output disk 22 and an upward force is applied by the input disk 21.As a result, both of the upper and the lower rollers 231, 232 incline sothat the contact points with the input disk 21 move to the inner side inthe radial direction and that with the output disk 22 moves to the outerside in the radial direction, and consequently the speed ratio of thecontinuously variable transmission mechanism 20 becomes larger (decreaseof speed). The supply operation of the hydraulic pressure for increasingand decreasing speed, PH, PL by the hydraulic pressure control section111 will be described in detail later in the description for thehydraulic pressure control circuit.

The structure and the operation of the first continuously variabletransmission mechanism 20 described above may be also applied to thesecond continuously variable transmission mechanism 30.

As shown in FIG. 2, the input disks 21, 31 of the first and the secondcontinuously variable transmission mechanisms 20, 30 are respectivelyspline-fitted to respective end portions of the primary shaft 12 havinga through-hole inside into which the input shaft 11 is inserted with aplay so that the input disks 21 and 31 always rotate with the samespeed, and since the output disks 22, 23 of the first and the secondcontinuously variable transmission mechanisms 20, 30 are integrated intoone unit as described previously, the rotation speeds of the outputsides of the first and the second continuously variable transmissionmechanisms 20, 30 are also kept always to be identical. To keep in stepwith these matters, the speed ratio control of the first and the secondcontinuously variable transmission mechanisms 20, 30 by the inclinationcontrol of the rollers 23, 23 is implemented so that said speed ratiomay be always kept to be identical.

As shown in the enlarged view of FIG. 4, the first gear 91, formed intoring shape, of the high mode gear train 90 is fitted onto the outersurface of the integrated output disk 34 and is fixed thereto bywelding, wherein, on the one side surface of the integrated output disk34, a circular groove Y is formed between the outer surface of said disk34 and the inner surface of the first gear 91, and the disk 34 and thegear 91 are weld-jointed in this groove Y.

Therefore, even if the weld metal Z stands up from the welding surface,this does not interfere with the toroidal surface 34 a formed on saidone side surface, so that the roller can be inclined and rotated in widerange. Further, since the first gear 91 is fixed onto the outer surfaceof the integrated output disk 34 by welding, an axial play of the firstgear 91 can be controlled and the support thereof is stabilized. On theother hand, as shown in FIGS. 5 and 6, the loading cam 40 has a cam disk41 interposed between the first gear 81 of the low mode gear 80 and theinput disk 21 of the first continuously variable transmission mechanism20, wherein the surfaces of the cam disk 41 and the input disk 21 whichare facing with each other are respectively formed into cam surfaceshaving convex and concave section continuously repeating in thecircumferential direction and a plurality of rollers held by a retainerdisk 42 is interposed between these cam surfaces.

The cam disk 41 is connected to the first gear 81 of the low mode geartrain 80, which is spline-fitted onto the input shaft 11 at the endthereof located on the opposite side with respect to the engine, by theplurality of pin members 44 disposed parallel to the shaft linedirection so as to be rotated together with it, and, as shown in FIG. 6,coned disk springs 45, 45, a needle thrust bearing 46 and a bearing race47 thereof are interposed between the cam disk 41 and a flange 12 aformed on the primary shaft 12, so that the cam disk 41 is pressed ontothe input disk 21 side by the spring force of the coned disk springs 45,45.

Thereby, the rollers 43, 43 are held between concaved portions 21 a, 41a on the cam surfaces of the disks 21, 41, and transmit the torque,which is inputted from the input shaft 11 through the first gear 81 ofthe low mode gear train 80 into the cam disk 41, to the input disk 21 ofthe first continuously variable transmission mechanism 20, and furthertransmit it through the primary shaft 12 to the input disk 31 of thesecond continuously variable transmission mechanism 30.

As especially shown by the chain line in FIG. 5, the rollers 43, 43 rollfrom the concaved portions 21 a, 41 a toward the convex portions 21 b,41 b on the cam surfaces of the disks 21, 41 in response to themagnitude of the input torque and are stuck between both cam surfaces,and thereby the input disk 21 of the first continuously variabletransmission mechanism 20, the roller 23, the integrated output disk 34,and the roller 33 of the second continuously variable transmissionmechanism 30 are pressed in this order toward and onto the input disk 31of the second continuously variable transmission mechanism 30.Accordingly, the holding pressure applied onto the roller 23, 33 of thefirst and the second continuously variable transmission mechanisms 20,30 are automatically adjusted in response to the magnitude of the inputtorque.

Further, in the loading cam 40, the pin members 44, . . . , 44 whichconnect cam disk 41 and the first gear 81 of the low mode gear train 80are disposed in the convex portions 41 b, . . . , 41 b of the cam disk41 where the thickness thereof is rather thicker. Accordingly, an axiallength of the cam disk 41 need not be increased improperly by makingoverall thickness thereof thicker, and the strength of the cam disk 41is prevented from being weakened by disposing insert holes of the pinmembers 44, . . . , 44 close to the concaved portions 41 a, . . . , 41 aformed on the cam surface.

To explain a support structure of the primary shaft 12 into which theinput shaft 11 is fittingly inserted with play with reference to FIG. 6,the engine side end of the primary shaft 12 is supported by thetransmission case 100 through a bearing 131, and onto the other endthereof is spline-fitted the first gear 81 of the low mode gear train80, and said gear 81 is supported through a bearing 132 by a cover 101located on the opposite side, with respect to the engine, of thetransmission case 100.

A coned disk spring 135 which applies force to the primary shaft 12 andthe first gear 81 in the direction to separate them with each otherthrough a needle thrust bearing 133 and a bearing race 134 is disposedbetween the first gear 81 and the flange 12 a which supports the coneddisk spring 45, 45 of the loading cam 40 on the primary shaft 12.

Accordingly, since, when the primary shaft 12 is expanded or contracteddue to the thermal expansion and the like, the engine side end of theshaft 12 is not allowed to move in the axial direction, the other endthereof which is spline-fitted into the first gear 81 makes displacementin the axial direction, and, at that time, said displacement is absorbedby the coned disk spring 135 and the first gear 81 is always pressedonto the bearing 132 side by an appropriate force in response to thespring force of the coned disk spring 135. Thereby, the condition wherethe first gear 81 is strongly pressed onto the bearing 132 side due tothe expansion of the primary shaft 12 or the first gear 81 has an axialplay due to the contraction of the primary shaft 12 can be avoided.

An appropriate force is always applied to the engine side and theopposite side bearings 131, 132 to which the spring force of the coneddisk spring 135 is applied through the primary shaft 12 and the firstgear 81, and thereby, even if the bearings 131, 132 are tapered-rollerthrust bearings as shown in the drawing, an axial pre-load can bemaintain properly so that the rattling or the increased rotationresistance problems caused by too small or too much pre-load can beavoided.

An oil pump 102 is mounted on the cover 101 located on the opposite sidewith respect to the engine, and is driven by the first gear 81 of thelow mode gear train 80 which rotates integrally with the input shaft 11.

Then the structure of the secondary shaft 13, and the planetary gearmechanism 50, the low mode clutch 60, and the high mode clutch 70, eachbeing mounted on the secondary shaft 13, will be described withreference to FIG. 7.

The secondary shaft 13 is rotatably supported at one end by a cover 103located in the engine side of the transmission case 100 and at the otherend by the cover 101 located in the opposite side thereof throughbearings 141, 142 respectively. The second gear 92 included in the highmode gear train 90 is disposed on the middle of the secondary shaft 13and the planetary gear mechanism 50 is disposed adjacent to the backside (this means the distal side with respect to the engine and will beused below in the same meaning) of the second gear 92, wherein thesecond gear 92 is coupled with the sun gear 52 of the planetary gearmechanism 50. In the back side thereof, a flange member 54 connected tothe internal gear 53 of the planetary gear mechanism 50 is spline-fittedonto the secondary shaft 13. Further, the low mode clutch 60 is disposedin the back side of the planetary gear mechanism 50. The clutch 60 isrotatably supported on the secondary shaft 13 and comprises a drummember 61 to which the secondary gear 82 of the low mode gear train 80is fixed, a hub member 62 which is disposed within the drum 61 in theradial direction and is connected to the pinion carrier 51 of theplanetary gear mechanism 50 through the flange member 55, a pluralityclutch plates 63, . . . , 63 each of which is alternately spline-fittedto each of said members 61, 62, and a piston 64 disposed in the drummember 61.

A hydraulic pressure chamber 65 is formed between the piston 64 and thedrum member 61 located in the back side thereof, and, when the hydraulicpressure for coupling generated by the clutch control unit 120 shown inFIG. 3 is introduced into the pressure chamber 65, the piston 64 ismoved to the fore side (this means the proximal side with respect to theengine and will be used below in the same meaning) while compressing aspring 66, and the clutch plates 63, . . . , 63 are engaged, and therebythe second gear 82 of the low mode gear train 80 is coupled with thepinion carrier 51 of the planetary gear mechanism 50 through the clutch60.

A balance piston 67 is disposed on the fore side of the piston 64, andthe pressure applied to the piston 64 by the centrifugal force appliedto the operating fluid in the hydraulic pressure chamber 65 is offset byintroducing lubricating oil into a balance chamber 68 formed betweenboth pistons 64, 67.

The high mode clutch 70 is disposed on the fore side of the second gear92 of the high mode gear train 90. The clutch 70 comprises a drum member71 which is connected through a gear 4 d for a parking mechanism to afirst gear 4 a of the output gear train 4 which is spline-fitted ontothe secondary shaft 13, a hub member 72 which is disposed within saiddrum member 71 in the radial direction and is connected to the secondgear 92, a plurality of clutch plates 73, . . . , 73 each of which isalternately spline-fitted to each of said members 71, 72, and a piston74 disposed in the drum member 71.

When the hydraulic pressure for coupling generated by the clutch controlunit 120 is introduced into a pressure chamber 75 formed on the backside of the piston 74, the piston 74 is moved to the back side whilecompressing a spring 76 and the clutch plates 73, . . . , 73 areengaged, and thereby the second gear 92 of the high mode gear train 90is coupled through the clutch 70 with the secondary shaft 13 and thefirst gear 4 a of the output gear train 4 spline-fitted to said shaft13. A balance piston 77 is also disposed on the back side of the piston74, and the pressure applied to the piston 74 by the centrifugal forceapplied to the operating fluid in the hydraulic pressure chamber 75 isoffset by introducing the lubricating oil into a balance chamber 78formed between both pistons 74, 77. Further, on the back side end of thesecondary shaft 13, a recessed portion 13 a extending toward fore sidefrom the end surface thereof is formed, and a boss 101 a which is formedon and extruded from the back side cover 101 is rotatably inserted intosaid recessed portion 13 a. A boss 103 a extruding toward back side isalso formed on the fore side cover 103 and is rotatably inserted into arecessed portion 13 b formed in the fore end of the secondary shaft 13.

Two oil channels 151, 161 for controlling the low mode and the high modeclutches 60, 70 are formed in the boss 101 a of the back side cover 101parallel to the axial line thereof, and oil channels 152, 162 which runupward from the clutch control unit 120 through the back side cover 101are connected to the oil channels 151, 161 respectively. Among these oilchannels, the oil channel 151 for the low mode clutch 60 communicateswith the hydraulic pressure chamber 65 of said clutch 60 through aradial hole 153 formed in the boss 101 a of the back side cover 101, acircumferential channel 154 formed on the outer surface of the boss 101a, a radial hole 155 formed in a circumferential wall of a recessedportion 13 a of the secondary shaft 13 into which the boss 101 a isinserted, a circumferential channel 156 formed on the outer surface ofthe shaft 13, and the through-hole 157 formed in a boss of the drummember 61 of the low mode clutch 60. Thereby, the hydraulic pressure forcontrolling the low mode clutch generated by the clutch control unit 120is introduced into the hydraulic pressure chamber 65 of the clutch 60.The oil channel 161 for the high mode clutch 70 is open at the front endof the boss 101 a and communicates with a space 163 between the frontend surface of said boss and the inner end surface of the recessedportion 13 a of the secondary shaft 13. Further, this oil channelcommunicates with an oil channel 164 which is formed by drilling on thesecondary shaft 13 along the axial direction thereof and whose back sideend is open to the inner end surface of the recessed portion 13 a, andfurther communicates with the hydraulic pressure chamber 75 of the highmode clutch 70 through radial holes 165, 166 formed in the secondaryshaft 13 and in the first gear 4 a of the output gear train 4respectively. Thereby, the hydraulic pressure for controlling the highmode clutch generated by the clutch control unit 120 is introduced intothe hydraulic pressure chamber 75 of the clutch 70.

Thus, since either of the oil channels 161, 161 for controlling the lowmode and the high mode clutches 60, 70 is introduced from the back sidecover 101 where the oil pump 102 is installed and respectivelycommunications with the hydraulic pressure chambers 65, 75 of theclutches 60, 70 that the secondary shaft 13, the hydraulic pressure canbe supplied more promptly to the hydraulic pressure chambers 65, 75comparing with, for example, the case where one of the oil channels isintroduced from the fore side cover 103, so that the coupling controlresponse of the clutches 60, 70 can be improved.

An oil channel 171 is formed in the boss 103 a of the fore side cover103 and is connected to an oil channel 172 (see FIG. 2) which runs fromthe clutch control unit 120 upward through the cover 103. Further, anoil channel 174 for the lubricating oil is drill-formed extending fromthe recessed portion 13 b formed in the fore end of the secondary shaft13, into which the boss 103 a is inserted, along the axial direction andis closed by a plug 173 at a back side end thereof, and a plurality ofradial through-holes 175, . . . , 175 which communications with the oilchannel 174 is formed in predetermined positions of the secondary shaft13. Thereby, the lubricating oil supplied from the clutch control unit120 can be supplied to the balance chambers 68, 78 of the low mode andthe high clutches 60, 70, and other lubrication points.

As shown in FIG. 3, the shift control unit 110 and the clutch controlunit 120 for controlling the low mode and the high mode clutches 60, 70are respectively mounted on the lateral side and on the bottom side ofthe transmission case 100, so that, by employing this structure wherethe control unit is divided into two units and one is mounted on thelateral side and the other is mounted on the bottom side of thetransmission case 100, the downward overhung length from thetransmission case can be made shorter comparing with the case where bothcontrol units are integrated into one unit and mounted on the bottom ofthe transmission case. Accordingly, this provides an advantage inkeeping a minimum road clearance.

As described above, since the shift control unit 110 is disposed on oneside (the left side in FIG. 3) of the transmission case 100, and thetrunnions 25, 25 are respectively attached to the upper and the lowerrods 27, 27 which extends horizontally toward inside of the transmissioncase 100 from the trunnion control section 112 of said unit 110 to movethe trunnions 25, 25 along the horizontal shaft center lines X, X, awide space is not required in the width direction, which is necessary inthe case where the trunnion is moved in the vertical direction andthereby the trunnion driving section is disposed on the top of thetransmission case.

Accordingly, when the secondary shaft 13 on which the planetary gearmechanism 50, the low mode and the high mode clutches 60, 70 are mountedis designed, the shaft center thereof can be placed near to the shaftcenters of the input shaft 11 and the primary shaft 12, so that thetransmission 10 can be made compact as a whole.

Supply control of the hydraulic pressure fluid for controlling the lowmode and the high mode clutches 60, 70 by the clutch control unit 120will be described in detail later in the item for the hydraulic pressurecontrol circuit.

Then, the mechanical operation of the continuously variable transmission10 having a structure described above will be explained.

When a vehicle is at a stop, in FIGS. 1 and 2, the transmission 10 is inthe low mode where the low mode clutch 60 is engaged and the high modeclutch 70 is disengaged, and the revolution from the engine 1 istransmitted from the back side end of the input shaft 11 through the lowmode gear train 80 composed of the first gear 81, the idle gear 83, andthe second gear 82 to the secondary shaft 13, and, at the same time, isinputted through the low mode clutch 60 into the pinion carrier 51 ofthe planetary gear mechanism 50.

The revolution inputted from the engine 1 into the input shaft 11 isfurther inputted from the first gear 81 of the low mode gear train 80through the loading cam 40 located adjacent thereto into the input disk21 of the first continuously variable transmission mechanism 20, andthen is transmitted through the rollers 23, 23 to the integrated outputdisk 34, and, at the same time, is also inputted from the input disk 21through the primary shaft 12 into the input disk 31 of the secondcontinuously variable transmission mechanism 30 disposed on the foreside end of said shaft 12, and then, same as the first continuouslyvariable transmission mechanism 20 does, is transmitted through therollers 33, 33 to the integrated output disk 34. At that time, theinclination angle of the roller 23, 33 of the first and the secondcontinuously variable transmission mechanisms 20, 30, that is, the speedratios of both continuously variable transmission mechanisms 20, 30, arekept to an identical predetermined value by the shift control unit 110shown in FIG. 3 through the control of the hydraulic pressure forincreasing speed PH and that for decreasing speed PL.

The rotation of the integrated output disk 34 of the first and thesecond continuously variable transmission mechanisms 20, 30 istransmitted through the high mode gear train 90 composed of the firstgear 91 formed on the outer surface of the integrated output disk 34 andthe second gear 92 on the secondary shaft 13 into the sun gear 52 of theplanetary gear mechanism 50.

Thus, the rotation is inputted into the planetary gear mechanism 50through the pinion carrier 51 and also through the sun gear 52, and, atthat time, by setting the rotation speed ratio between the pinioncarrier 51 and the sun gear 52 to a predetermined ratio by the speedratio control of the first and the second continuously variabletransmission mechanisms 20, 30, the rotation of the internal gear 53 ofthe planetary gear mechanism 50, that is, the rotation inputted from thesecondary shaft 13 through the output gear train 4 into the differentialgear unit 6 is set to zero to make the geared neutral condition of thetransmission 10.

Starting from this condition, when the ratio of the input rotationspeeds of the pinion carrier 51 and the sun gear 52 is changed bychanging the speed ratio of the first and the second continuouslyvariable transmission mechanisms 20, 30, the internal gear 53 or thesecondary shaft 13 is rotated in the forward or the reverse directionunder a condition where the speed ratio as a whole transmission 10(hereafter referred to as “final speed ratio”) is rather high, that is,under low mode condition, and the vehicle starts to move.

After the vehicle starts to move forward, when the low mode clutch 60 isdisengaged and the high mode clutch 70 is engaged at a predeterminedtiming, the revolution inputted from the engine 1 into the input shaft11 is transmitted, same as the low mode does, from the loading cam 40 tothe input disks 21, 31 of the first and the second continuously variabletransmission mechanisms 20, 30, and is further transmitted throughrespective roller 23, 33 to the integrated output disk 34, and, at thesame time, is transmitted through the high mode gear train 90 and thehigh mode clutch 70 to the second shaft 13.

At that time, the planetary gear mechanism 50 is in a racing conditionand the final speed ratio is determined only by the speed ratios of thefirst and the second continuously variable transmission mechanisms 20,30 so that the final speed ratio is continuously controlled under highmode condition where the final speed ratio is small.

According to this transmission 10, since the low mode gear train 80which transmits the rotation from the input shaft 11 to the planetarygear mechanism 50 on the secondary shaft 13 under the geared neutral orthe low mode condition is disposed on the back side of the input shaft11 and the secondary shaft 13, the gear train 80 does not interfere withthe differential gear unit 5 which is disposed on the fore side end ofthe secondary shaft 13 and the output gear train 4 which transmits thepower to the differential gear unit 5, and thereby the axial lengthelongation of the transmission 10, which may occur when the gear trainis offset in the axial direction to prevent this interference, can beavoided.

When, as shown in this transmission 10, the first and the secondcontinuously variable transmission mechanisms 20, 30 are employed as thetoroidal type continuously variable transmission mechanism, and theinput disks 21, 31 thereof are coupled with both ends of the primaryshaft 12 respectively, and the output disks 22, 32 are disposed on themiddle of the primary shaft 12, and the low mode gear train 80 whichtransmits the rotation to the secondary shaft 13 side is disposed on theback side end of the input shaft 11, the arrangement of the loading cam40 which is interposed between the input shaft 11 and the input portionto the first and the second continuously variable transmissionmechanisms 20, 30, that is, where the loading cam 40 shall be placed,should be carefully examined.

When, as shown in FIG. 8, the loading cam 40′ is placed between theinput shaft 11′ and the input disk 31′ of the continuously variablemechanism 30′ located in the engine 1′ side, in the low mode, the torquefrom the engine 1′ is transmitted, as shown by an arrow “a”, from theback side end of the input shaft 11′ through the gear train 80′ into thesecondary shaft 13′ side, and, at the same time, a reaction torquegenerated in the planetary gear mechanism 60′ on the secondary shaft 13′is circulated back, as shown by an arrow “b”, to the output disk 34′ ofthe continuously variable mechanisms 20′, 30′ through the gear train 90′to make a circulating torque, and after being transmitted to the inputdisks 21′, 31′ of the continuously variable mechanisms 20′, 30′, thiscirculating torque is inputted from the input disk 31′ of thecontinuously variable mechanism 30′ in the engine side through theloading cam 40′ into the input shaft 11′ again and is transmittedfurther to the back side gear train 80′ again.

Thus, the torque from the engine 1′ (arrow “a”) and the circulatingtorque (arrow “b”) are parallelly applied to the input shaft 11′, sothat the shaft 11′ should be reinforced by, for example, increasing thediameter thereof. As a result, the total weight of the transmission 10is increased and, since the rigidly of the input shaft 11′ is increasedand thereby the vibration of the engine 1′ likely to be transmitted tothe output side, the vibration and the noise of the vehicle isincreased.

On the contrary, according to the continuously variable transmission 10of the preferred embodiment, since the low mode gear train 80 whichtransmits the rotation to the secondary shaft 13 side is disposed on theback side end of the input shaft 11 and the loading cam 40 which isinterposed between the input shaft 11 and the continuously variabletransmission mechanisms 20, 30 is also disposed on the back side end ofthe input shaft 11, the strength and rigidity problem of the input shaft11 described above can be avoided.

In this case, as shown in FIG. 9, while the torque from the engine 1 istransmitted, as shown by an arrow “c”, from the back side end of theinput shaft 11 through the low mode gear train 80 to the secondary shaft13 side, the circulating torque from the planetary gear mechanism 50 onthe secondary shaft 13 is transmitted, as shown by an arrow “d”, throughthe high mode gear train 90, back to the output disk 34 of thecontinuously variable transmission mechanisms 20, 30, and then thiscirculating torque is respectively transmitted, in the firstcontinuously variable transmission mechanism 20, from the input disk 21through the loading cam 40 to the first gear 81 of the low mode gear 80directly and, in the second continuously variable transmission mechanism30, from the input disk 31 through the primary shaft 12 and the sameloading cam 40 to the first gear 81 of the low mode gear 80. Thus, eachcirculating torque transmitted back to the first and the secondcontinuously variable transmission mechanisms 20, 30 does not passthrough the input shaft 11, so that the input shaft 11 is allowed totransmit only the torque from the engine 1. As a result, the input shaft11 may have small diameter and the transmission 10 may have lightweight, and, at the same time, decreased rigidity of the input shaft 11allows to effectively absorb the vibration of the engine 1 and therebythe vibration and noise of the vehicle is made lower.

Hydraulic Pressure Control Circuit

There will not be describe the hydraulic pressure control circuit of thecontinuously variable transmission 10 comprising the shift control unit110 and the clutch control unit 120.

As shown in FIG. 10, a hydraulic pressure control circuit 200 comprisesa regulator valve 202 which regulates a pressure of an operating fluiddischarged from an oil pump 102 into a predetermined line pressure andoutputs to a main line 201, a relief valve 204 which generates apredetermined relief pressure taking the line pressure supplied throughthe main line 201 as a source pressure and outputs said relief pressureto a relief line 203, and a manual valve 208 which is operated by therange shifting operation of the vehicle driver, and makes the main line201 communicates with a first and a second output lines 205, 206 in theD-range and with the first and a third output lines 205, 207 in theR-range, and intercepts the line pressure in the N-range and theP-range. The regulator valve 202 and the relief valve 204 are equippedwith a linear solenoid valve 209 for the line pressure and a linearsolenoid valve 210 for the relief pressure respectively, and further areducing valve 211 which makes a constant pressure taking the linepressure as a source pressure is provided, wherein the linear solenoidvalves 209, 210 make control pressures respectively base on the constantpressure generated by the reducing valve 211. By supplying the controlpressures to the control ports 202 a, 204 a of the regulator valve 202and the relief valve 204, pressure regulation values of the linepressure and the relief pressure are controlled by respective linearsolenoid valves 209, 210.

The hydraulic pressure control circuit 200 further comprises athree-layers valve for forward running 220 and a three-layers valve forbackward running 230 which respectively generate the hydraulic pressurefor increasing speed PH and the hydraulic pressure for decreasing speedPL for implementing shift control in forward and backward running basedon the line pressure and the relief pressure, and a shift valve 241which selectively actuates these three-layers valves 220, 230.

The position of a spool of the shift valve 241 is determined by the linepressure supplied as a control pressure into a control port 241 a formedon one end thereof, that is, when the line pressure is not supplied, thespool is in the right side so that the main line 201 communicates with aline pressure supply line 242 which leads to the three-layers valve forforward running 220, and, when the line pressure is supplied, the spoolis in the left side so that the main line 201 communicates with a linepressure supply line 243 which leads to the three-layers valve forbackward running 230. The three-layers valves for forward and backwardrunning 220, 230 have the same structure, and each has a sleeve 222, 232which is axial-movably fitted into a bore 221, 231 (see FIG. 11) formedin a valve body 111 a of a hydraulic pressure control section 111 of theshift control unit 110 shown in FIG. 3 and a spool 223, 233 which isalso axial-movably fitted into said sleeve 222, 232. A line pressureport 224, 234 to which is connected the line pressure supply line 242,243 led from the shift valve 241 is provided on the middle portionthereof, and a first and a second relief pressure ports 225, 226, 235,236 to which are connected lines branched from the relief pressure line203 are provided on both end portions thereof respectively, and anacceleration pressure port 227, 237 is provided between the linepressure port 224, 234 and the first relief pressure port 225, 235, anda deceleration pressure port 228, 238 is provided between the linepressure port 224, 234 and the second relief pressure port 226, 236respectively. The operation of this three-layers valves 220, 230 will beexplained by taking the three-layers valves for forward running 220 asan example. When the relative position between the sleeve 222 and thespool 223 is changed from the neutral condition, which is shown in FIG.10, to the position where the sleeve 222 is relatively moved to theright on the drawing, a communication level between the line pressureport 224 and the acceleration pressure port 227 as well as acommunication level between the second relief pressure port 226 and thedeceleration pressure port 228 are respectively increased, and, on thecontrary, when the sleeve 222 is relatively moved to the left, acommunication level between the line pressure port 224 and thedeceleration pressure port 228 as well as a communication level betweenthe first relief pressure port 225 and the acceleration pressure port227 are respectively increased.

A line 244, 245 respectively led from the acceleration pressure port227, 237 of the three-layers valve for forward or backward running 220,230 and a line 246, 247 respectively led from the deceleration pressureport 228, 238 of the three-layers valve for forward or backward running220, 230 are connected to the shift valve 241.

When the spool of the shift valve 241 is in the right, the lines 244,246 led from the acceleration pressure port 227 and the decelerationpressure port 228 of the three-layers valve for forward running 220communicate respectively with the acceleration pressure line 248 and thedeceleration pressure line 249 which respectively communicate with thehydraulic pressure chambers for increasing speed 1151, 1152 and thehydraulic pressure chambers for decreasing speed 1161, 1162 of thetrunnion control section 112 of the shift control unit 110 shown in FIG.3, and, on the contrary, when the spool of the shift valve 214 is in theleft, the lines 245, 247 led from the acceleration pressure port 237 andthe deceleration pressure port 238 of the three-layers valve forbackward running 230 communicate respectively with the accelerationpressure line 248 and the deceleration pressure line 249 whichrespectively communicate with the hydraulic pressure chambers forincreasing speed 1151, 1152 and the hydraulic pressure chambers fordecreasing speed 1161, 1162.

As shown in FIG. 11, the sleeve 222, 232 of the three-layers valve forforward or backward running 220, 230 is axially driven by a step motor251, 251 respectively. Further, a cam mechanism 260 which moves thespool 223, 233 in an axial direction against spring force generated by aspring 229, 239 in response to the movement of the sleeve 222, 232 bythe step motor 251, 251 is provided.

As shown in FIGS. 11 and 12, the cam mechanism 260 comprises aprecession cam 261 which has a cam surface 261 a formed into a helicalsurface on one end and is attached to a predetermined trunnion, moreparticularly, to an end of a rod 37 of a first trunnion 351 located onthe upper portion of the second continuously variable transmissionmechanism 30, a shaft 262 which is disposed on one end of the spool 223,233 of the three-layers valve for forward or backward running 220, 230in the direction normal thereto and is rotatably supported by the valvebody 111 a of the hydraulic pressure control section 111, a driven lever263 which is attached to one end of the shaft 262 and is brought intocontact with the cam surface 261 a on a free end thereof, and drivelevers for forward or backward running 264, 265 which are also attachedto the shaft 262 and whose free ends are engaged with slits 223 a, 233 aformed on ends of the spools 223, 233 of the three-layers valve forforward or backward running 220, 230 respectively.

When the first trunnion 351 and the rod 37 is integrally rotated aroundthe shaft center line X by the inclination of the first roller 331 ofthe second continuously variable transmission mechanism 30, theprecession cam 261 is also rotated together with them and the drivenlever 263 whose free end if brought into contact with the cam surface261 a of the precession cam 261 is swung by a predetermined angle and,through the shaft 262, the drive levers for forward or backward running264, 265 are also swung by the same angle, and thereby the spools 223,233 of the three-layers valve for forward or backward running 220, 230are axially moved respectively by a stroke corresponding to the swingangle of the drive levers 264, 265.

Accordingly, the position of the spools 223, 233 always corresponds tothe inclination angle of the roller 33 of the second continuouslyvariable transmission mechanism 30 (and the roller 23 of the firstcontinuously variable transmission mechanism 20), in other words, to thespeed ratio of these continuously variable transmission mechanisms 20,30.

According to the cam mechanism 260, since the spools 223, 233 of thethree-layers valve for forward or backward running 220, 230 are drivenby the single precession cam 261 and the driven lever 263 as describedabove, the structure of the cam mechanism is simplified comparing withthe case having respective precession cams for the spools 223, 233.

Since the step motor 251, 252 is directly attached to the side surfaceof the valve body 111 a of the hydraulic pressure control unit 111 ofthe shift control unit 110, in which the three-layers valves for forwardand backward running 220, 230 are built, with an axial center line incommon with the corresponding three-layers valve 220, 230 and isdirectly coupled with the sleeve 222, 232 of the three-layers valve 220,230 respectively, the mechanism for driving the sleeve 222, 232 of thethree-layers valve 220, 230 by the step motor 251, 252 is considerablysimplified comparing with the case where the step motor is disposedindependently from the three-layers valve, for example, on the covermember of the transmission case, the oil pan and the like, and both areconnected with each other through some coupling mechanism, and, inaddition, the position of the sleeve 222, 232 can be accuratelycontrolled.

Further, since, in this shift control unit 110, the shift valve 241 isdisposed between both three-layers valves for forward or backwardrunning 220, 230, oil channels between the shift valve 241 and boththree-layers valves 220, 230, in concretely, lines 242-247 in thehydraulic pressure control circuit shown in FIG. 10 can be made shorter,and thereby the control response using these three-layers valves 220,230 can be improved. On the other hand, the hydraulic pressure controlcircuit 200 is provided with a first and a second solenoid valves 271,272 for controlling the clutch, and the first and the second outputlines 205, 206 led from the manual valve 208 are connected to the firstand the second solenoid valves 271, 272 respectively.

When the first solenoid valve 271 is opened, a clutch engaging pressurebased on the line pressure from the first output line 205 is suppliedthrough a fall safe valve 273 and a low mode clutch line 274 to thehydraulic pressure chamber 65 of the low mode clutch 60 to makeengagement of the clutch 60, and when the second solenoid valve 272 isopened, the clutch engaging pressure base on the line pressure from thesecond output line 206 is supplied through a high mode clutch line 275to the hydraulic pressure chamber 75 of the high mode clutch 70 to makeengagement of the clutch 70.

The low mode and the high mode clutch lines 274, 275 are provided withaccumulators 276, 277 respectively to gently supply the engagingpressure to the low mode and the high mode clutches 60, 70 and therebyto eliminate the shock during engagement.

The third output line 207 led out from the manual valve 208 is connectedthrough the fail safe valve 273 to the control port 241 a of the shiftvalve 241, and, when the manual valve 208 is shifted to the position ofR-range, the line pressure is supplied to the control port 241 a of theshift valve 241 to move the spool of the shift valve 241 to the left,that is, to the position for backward running.

Further, a fail safe solenoid valve 278 is provided for actuating thefail safe valve 273, and when the spool of the fail safe valve 273 isshifted to the right by the control pressure from the solenoid valve278, the first output line 205 is made to communicate with the low modeclutch line 274.

All of the first and the second solenoid valves 271, 272 and the failsafe solenoid valve 278 are of three-way valves, and, when the line isintercepted by the valve into the upstream and the downstream sides, thedownstream line is drained.

The clutch control unit 120, in which the first and the second solenoidvalves 271, 272 and the like are disposed, comprises, as shown in FIG.13, an upper member 121, a middle member 122, and a lower member 123connected into one unit by a plurality of bolts 124, . . . , 124, andthe first and the second solenoid valves 271, 272 are attached onto aside surface of the middle member 122 by a fixing plate 125.

At that time, the solenoid valves 271, 272 are fixed by placing flanges271 a, 272 a formed on the outer surfaces of main bodies of the solenoidvalves 271, 272 between the fixing plate 125 and the side surface of themiddle member 122, and the fixing plate 125 is fastened to the upper andthe lower members 121, 123 by bolts 126, 126, that means, the uppermember 121 and the lower member 123 are connected with each otherthrough the fixing plate 125, and thereby the rigidity of the clutchcontrol unit 120 constructed as a three-layers structure is improved asa whole.

Adding to the structure described above, the hydraulic pressure controlunit 200 shown in FIG. 10 is provided with a lubrication line 281. Thelubrication line 281 is led from the drain port of the regulator valve202 and is branched into a line 282 for supplying the lubricating oil toeach lubrication point in the first and the second continuously variabletransmission mechanism 20, 30 of the transmission 10, and a line 283 forsupplying the lubricating oil to the portions other than thecontinuously variable transmission mechanisms 20, 30 such as theplanetary gear mechanism 50, balance chambers 68, 78 of the low mode andthe high mode clutches 60, 70, and the like, and a relief valve 284 forcontrolling a pressure of the lubricating oil to be a predeterminedvalue is connected to the line 281.

An upstream portion of the line 282 communicating with the continuouslyvariable transmission mechanisms 20, 30 is branched into a cooling line286 on which a cooler 285 for cooling the lubrication oil is providedand a bypass line 287 for bypassing the cooler 285, and an orifice 288and a first switching valve 289 are parallelly disposed on the coolingline 286 at the upper stream side of the cooler 285, and a secondswitching valve 290 for opening or closing the bypass line 287 isdisposed on said line 287.

A supply control of the lubricating oil to the continuously variabletransmission mechanisms 20, 30 by the first and the second switchingvalve 289, 290 will be described below.

First, the second switching valve 290 is opened by a signal from acontrol unit 300 (see FIG. 14), which will be described later, when thetemperature of the operating fluid is lower than the predetermined valueor when the hydraulic pressure of the operating fluid is higher than thepredetermined value, and the lubricating oil is supplied to thecontinuously variable transmission mechanisms 20, 30 without passingthrough the cooler 285. This is because the lubricating oil need not becooled by the cooler 285 when the oil temperature is lower and shall besupplied effectively through the bypass line 287 having smallerresistance, and because the damage and the deterioration in durabilityon the cooler 285 which might be caused when the high pressure oilpasses through the cooler 285 shall be avoided.

In the case other than that described above, the second switching valve290 is closed and the lubricating oil is supplied to the continuouslyvariable transmission mechanisms 20, 30 after being cooled by the cooler285, and thereby the oil film of the lubricating oil on the toroidalsurfaces of the output disks 21, 22, 31, 32 is maintained to be propercondition and the durability of the toroidal surface and the surface ofthe rollers 23, 33 contacting therewith can be secured. The firstswitching valve 289 is controlled to be closed by the signal from thecontrol unit 300 when the second switching valve 290 is closed and therevolution speed of the engine 1 is lower than the predetermined valueor the vehicle speed is lower than the predetermined speed. This isbecause the clutches 60, 70 require a certain amount of lubricating oil,while the continuously variable transmission mechanisms 20, 30 requiresmaller amount thereof during low speed or low revolution driving, sothat lubricating oil supply to the continuously variable transmissionmechanisms 20, 30 shall be limited to secure that for the clutches 60,70 under these conditions where the lubricating oil is not suppliedsufficiently in the volume.

The lubricating oil supplied to the continuously variable transmissionmechanisms 20, 30 through the line 282 is supplied to the bearings ofthe rollers 23, 33 through the oil channel 282 a as shown in FIG. 3 andin injected also onto the toroidal surface of the output disks 21, 22,31, 32 by a nozzle 282 b.

(1) Basic operation of control

The continuously variable transmission 10 according to this embodimenthas a mechanical structure and the hydraulic pressure control circuit200 as described above, and also has a control unit 300 which implementsthe shift control of the transmission 10 as a whole by making a speedration control of the first and the second continuously variabletransmission mechanisms 20, 30 and an engage and disengage control ofthe clutches 60, 70 by the use of the hydraulic pressure control circuit200.

To the control unit 300 are inputted signals, as shown in FIG. 14, froma vehicle speed sensor 301 for sensing a vehicle speed, an engine speedsensor 302 for sensing an engine 1 speed, a throttle angle sensor 303for sensing a throttle angle of the engine 1, a range sensor 304 fordetecting a range selected by the driver, and further, for variouscontrol, from an oil temperature sensor 305 for sensing a temperature ofthe operating fluid, an input and an output rotation speed sensors, 306,307 for respectively sensing the input and the output rotation speeds ofthe continuously variable transmission mechanisms 20, 30, and idleswitch 308 for detecting a release of an accelerator pedal, a brakeswitch 309 for detecting a depressing on a brake pedal, an incline anglesensor 310 for sensing an incline angle of the road surface, and thelike.

The control signal is outputted to the linear solenoid valves 209, 210,for controlling the line and the relief pressures, the first and thesecond solenoid valves 271, 272 for the low mode and the high modeclutches 60, 70, the fail safe solenoid valve 278, the first and thesecond switching valves 289, 290 for the lubrication control, the stepmotors 251, 252 for the three-layers valves for forward and backwardrunning 220, 230 and the like, in response to the driving condition ofthe engine and the vehicle indicated by these sensors and the switches.

Then, the basic operation of the shift control by the hydraulic pressurecontrol circuit 200 and the control unit 300 will be described. In thedescription below, if not specified otherwise, the manual valve 208shown in FIG. 10 is in the D-range position and thereby the spool of theshift valve 241 is in the forward running position, which correspondsthe right position on the drawing, and, as to the continuously variabletransmission mechanism, the first roller 231 and the first trunnion 251located upper side of the first continuously variable transmissionmechanism 20 will be taken as an example for the explanation.

As for the speed ration control of the continuously variabletransmission mechanisms 20, 30 by the hydraulic pressure control circuit200, the linear solenoid valves 209, 210 for the regulator or the reliefvalves in the hydraulic pressure control circuit 200 are actuated togenerate the control pressures for the line pressure control and therelief pressure control respectively based on the signal from thecontrol unit 300, and these control pressures are respectively suppliedto the control ports 202 a, 204 a of the regulator and the relief valves202, 204 to generate the predetermined line pressure and thepredetermined relief pressure respectively.

Among these hydraulic pressure, the line pressure is supplied from themain line 201 through the shift valves 241, and the line 242 to the linepressure port 224 of the three-layers valve for forward running 220(hereafter referred to as “three-layers valve”). The relief pressure issupplied through the line 203 to the first and the second reliefpressure ports 225, 226 of the three-layers valve 220.

A pressure difference ΔP (=PH−PL) between the hydraulic pressure forincreasing speed PH and the hydraulic pressure for decreasing speed PLwhich are respectively supplied to the hydraulic pressure chamber forincreasing speed 115 (this means 1151, 1152, and will be used below inthe same meaning) and the hydraulic pressure chamber for decreasingspeed 116 of the shift control unit 110 by the three-layers valve 220 iscontrolled based on the line pressure and the relief pressure. Theobject of the pressure difference control is to hold the trunnion 25 orthe roller 23 in the predetermined neutral position against the tractionforce T applied to the trunnion 25 of the continuously variabletransmission mechanism 20, and to change the speed ratio of thecontinuously variable transmission mechanism 20 by moving the trunnion25 and the roller 23 from the neutral position along the axial centerline X and thereby inclining the roller 23.

As for the traction force T, as shown in FIG. 15, in the continuouslyvariable transmission mechanism 20, when the roller 23 is driven by therotation of the input disk 21 in “e” direction, to the roller 23 and thetrunnion 25 supporting said roller is applied the force for draggingthen in the same direction as the rotating direction “e” of the inputdisk 21. When the output disk 22 is driven into “g” direction (“x”direction in FIG. 3) by the rotation of the roller 23 in “f” direction,the force in the direction opposite to that of the rotation “g” of theoutput disk 22 is applied to the roller 23 and the trunnion 25 as areaction force. As a result, the traction force T with the directionshown in the drawing is applied to the roller 23 and the trunnion 25.

Accordingly, in order to hold the roller 23 in the neutral positionagainst the traction force T, the hydraulic pressures for increasing andfor decreasing speed PH, PL are respectively supplied to the hydraulicpressure chambers for increasing and decreasing speed 115, 116 which areformed by the pistons 113, 114 attached to the trunnion 25 through therod 27 so that the pressure difference ΔP balances with the tractionforce T.

When, for example, in order to decrease the speed ration of thecontinuously variable transmission mechanism 20 from this condition(acceleration), the sleeve 222 of the three-layers valve 220 is moved tothe left in FIG. 11 (to the right in FIG. 10), the communication levelbetween the line pressure port 224 and the acceleration pressure port227 and that between the second relief pressure port 226 and thedeceleration pressure port 228 of the three-layers valve 220 rise up.

Thereby, the hydraulic pressure for increasing speed PH supplied fromthe acceleration pressure line 248 shown in FIG. 10 to the hydraulicpressure chambers for increasing speed 115 is intensified by therelatively higher line pressure, and the hydraulic pressure fordecreasing speed PL supplied from the deceleration pressure line 249 tothe hydraulic pressure chamber for decreasing speed 116 is reduced bythe relatively lower relief pressure, and consequently the pressuredifference ΔP rise up, and, as a result, the pressure difference ΔPovercomes the traction force T and the trunnion 25 and the roller 23 aremoved to “h” direction shown in FIG. 15. This movement makes the roller23 inclined into the direction where the contact point with the inputdisk 21 moves radially outside and that with the output disk 22 movesradially inside, and thereby the speed ration of the continuouslyvariable transmission mechanism 20 is shifted to the acceleration side.

The inclination of the roller 23 occurs in the same manner in the secondcontinuously variable transmission mechanism 30 shown in FIG. 12, and amovement of the trunnion 35 in “i” direction caused by the pressuredifference ΔP superior to the traction force T makes the roller 33inclined into the direction where the contact point with the input disk31 moves radially outside and that with the output disk 32 movesradially inside, and the precession cam 261 of the cam mechanism 260rotates integrally with the inclination motion in the same direction(“j” direction in FIG. 11) by the same angle, and thereby all of thedriven lever 263, the shaft 262 and the drive lever 264 of the cammechanism 260 are rotated in “k” direction in FIG. 12.

As a result, the spool 223 of the three-layers valve 220 moves to “i”direction, to the left in FIG. 11, by the spring force of the spring229, and since this direction corresponds to that of the sleeve 222moved by the step motor 251, the communication level between the linepressure port 224 and the acceleration pressure port 227 and thatbetween the second relief pressure port 226 and the decelerationpressure port 228, which has been once risen up, is restored to aninitial neutral condition.

Thereby, the pressure difference ΔP is made to balance with the tractionforce again and the shift operation is completed, wherein the speedration of the continuously variable transmission mechanism 20 (and 30)is fix to new value with a certain amount of change.

At that time, this shift actuation finishes when the spool 223 moves tothe predetermined neutral position relative to the sleeve 222, and,since this position corresponds to that of the sleeve 222 moved by thestep motor 251 and also that determined by the incline angle of theroller 23 and the trunnion 25 through the cam mechanism 260, theposition of the sleeve 222 corresponds to the incline angle of theroller 23 and the trunnion 25. As a result, a controlled amount by thestep motor 251 corresponds to the speed ratio of the continuouslyvariable transmission mechanism 20 is controlled by the pulse control ofthe step motor 251 (and this can be applied also to the continuouslyvariable transmission mechanism 30).

Above actuation is implemented in the same manner when the sleeve 222 ofthe three-layers valve 220 is moved by the step motor 251 to theopposite direction, and, at that time, the speed ratio of thecontinuously variable transmission mechanism 20 is made larger (,wherethe car is decelerated). The characteristic of the change in the speedratio of the continuously variable transmission mechanisms 20, 30 withrespect to the number of the pulse of the control signal inputted intothe step motors 251, 252 is, for example, shown in FIG. 16, wherein thespeed ration becomes smaller (that is, the car is accelerated) as thenumber of the pulse increases.

Then, the control of the speed ration of the whole transmission 10(final speed ratio), which employs the speed ration control of thecontinuously variable transmission mechanisms 20, 30 described above,will be explained.

As described above, the speed ratio of the continuously variabletransmission mechanisms 20, 30 is controlled through the step control ofthe step motors 251, 252, wherein different final speed ratio is broughtdepending on whether transmission 10 is in the low mode or in the highmode, that is, which one of the low mode clutch 60 and the high modeclutch 70 is engaged.

In high mode, since to output rotation of the continuously variabletransmission mechanisms 20, 30 is directly transmitted to the secondaryshaft 13 through the high mode gear train 90 and the high mode clutch70, not through the planetary gear mechanism 50, as described above, thecharacteristic H of the final speed ratio with respect to the pulsenumber is, as shown in FIG. 17, similar to that of the speed ratio ofthe continuously variable transmission mechanisms 20, 30 shown in FIG.16. It is needless to say that the speed ratio values may differ witheach other depending on the difference in the diameter or the toothnumber of the first gear 91 and the second gear 92 of the high mode geartrain 90.

On the other hand, in the low mode, the revolution of the engine 1 isinputted from the input shaft 11 through the low mode gear train 80 andthe low mode clutch 60 into the pinion carrier 51 of the planetary gearmechanism 50, and, at the same time, the output rotation of thecontinuously variable transmission mechanisms 20, 30 is inputted throughthe high mode gear train 90 into the sun gear 52 of the planetary gearmechanism 50. At that time, when the ratio between the rotation speedinputted into the pinion carrier 51 and that inputted into the sun gear52 is set to a certain predetermined value by controlling the speedratio of the continuously variable transmission mechanisms 20, 30, therotation speed of the internal gear 53, which is an output element ofthe planetary gear mechanism 50, may become zero, that is, the gearedneutral condition may be obtained.

Under this condition, the final speed ratio becomes infinite as shown inFIG. 17 by the symbols “a”, “b”, and, when the speed ratio of thecontinuously variable transmission mechanisms 20, 30 is changed to thelarger side (deceleration side) to lower the input rotation speed intothe sun gear 52 by decreasing the pulse number of the control signal forthe step motors 251, 252 starting from this condition, the internal gear53 of the planetary gear mechanism 50 begins to rotate in the forwardrunning direction and the characteristic L in which the final speedration becomes smaller as the pulse number decreases is obtained, thatis, the low mode of the D-range is obtained. The curves of the low modecharacteristic L and the high mode characteristic H crosses with eachother at a predetermined pulse number (approximately 500 pulse in thedrawing) or at a predetermined speed ratio of the continuously variabletransmission mechanisms 20, 30 ( approximately 1.8 in the drawing),which is shown by “c” in the drawing. Therefore, when the low modeclutch 60 and the high mode clutch 70 are switched in this cross point“c”, the modes can be switched with continuously changing final speedratio.

When the speed ratio of the continuously variable transmissionmechanisms 20, 30 is changed to the smaller side (acceleration side) toraise the input rotation speed into the sun gear 52 by increasing thepulse number of the control signal for the step motors 251, 252 staringfrom the geared neutral condition, the internal gear 53 of the planetarygear mechanism 50 begins to rotate in the backward running direction andthe characteristic R of the R-range in which the final speed ratiobecomes larger as the pulse number increases is obtained.

Based on the control characteristics described above, the control unit300 controls the final speed ratio in response to the driving conditionof the vehicle.

The control unit 300 finds a current vehicle speed V and a throttleangle θ based on the signals from the vehicle speed sensor 301 and thethrottle angle sensor 303 and sets a target engine speed Neo by the useof these values and a predetermined map shown in FIG. 18. Then, in orderto obtain the corresponding final speed ratio to the target engine speedNeo (the value which corresponds to the angle α in FIG. 18), the controlunit 300 implements, based on the control characteristic shown in FIG.17, the pulse control for the step motors 251, 252 and the engagementcontrol of the low mode and the high mode clutches 60, 70 through thecontrol of the first and the second solenoid valves.

Adding to the speed ratio control of the continuously variabletransmission mechanisms 20, 30 by the pulse control of the step motors251, 252 (hereafter, referred to as “three-layers valve control”), thecontrol unit 300 of the transmission 10 also implements the speed ratiocontrol of the continuously variable transmission mechanisms 20, 30 bydirectly generating the predetermined pressure difference ΔP bycontrolling the relief pressure with the linear solenoid valve 210(hereafter, referred to as “direct control”). The reason thereof is asfollows.

Though the three-layers valve control is implemented on condition thatthere is a certain relation between the pulse number of the step motors251, 261 or the travel of the sleeves 222, 223 and the pressuredifference ΔP generated thereby, there might occur a hysteresis in thisrelation, for example, by the friction applied to the sleeves 222, 232during travelling as shown in FIG. 19, which shows different paths ofthe relation between the travelling of the sleeve in the increasingdirection of the pressure difference ΔP and that in the decreasingdirection. Thereby, there might occur an inversion of the pressuredifference ΔP placing a geared neutral position therebetween in a pointshown by “d” near the geared neutral (GN) due to the hysteresis, and, asa result, the driving direction might be inverted between the forwardand the backward running.

To cope with this problem, the pressure difference ΔP may be directlygenerated to supply to the hydraulic pressure chambers for increasing ordecreasing speed 115, 116, and the line pressure may be controlledtherefor, but the line pressure has a rather wider control range such as4-16 kg, so that it has a disadvantage in making a minute control of thepressure difference ΔP and also has another disadvantage that thehydraulic line pressure must be raised to make the predeterminedpressure difference ΔP resulting in a high pressure in the whole circuitand thereby an increase of the oil pump loss.

Therefore, when the pressure difference ΔP is generated, the reliefpressure which is lower than that of the line pressure has an advantagein making the pressure difference ΔP by lowering itself, and, because ofthe narrower control range of the relief pressure such as 0-4 kg, it canbe preferably used in minute control of the pressure difference ΔP.

In the direct control, the line pressure and the relief pressure aresupplied without being regulated by the three-layers valve 220 as thehydraulic pressures for increasing or decreasing speed PH, PL which aresupplied to the hydraulic pressure chambers for increasing or decreasingspeed 115, 116. When the sleeve 222 and the spool 223 of thethree-layers valve 220 are actuated from the neutral position shown inFIG. 10 to make the speed ration of the continuously variabletransmission mechanism 20 lower (acceleration), the sleeve 222 is, atfirst, moved to the right in the drawing by a predetermined stroke tomake the communication level between the line pressure port 224 and theacceleration pressure port 227 and that between the second reliefpressure port 226 and the deceleration pressure port 228 are raised upso that the line pressure is supplied from the acceleration pressureline 248 to the hydraulic pressure chamber for increasing speed 115 andthe relief pressure is supplied from the deceleration pressure line 249to the hydraulic pressure chamber for decreasing speed 116.

As a result, the trunnion 25 or the roller 23 are moved by the pressuredifference ΔP between the line pressure as a hydraulic pressure forincreasing speed PH and the relief pressure as a hydraulic pressure fordecreasing speed PL to the acceleration direction to incline the roller23, and the spool 223 is moved by the cam mechanism 260 to the samedirection with the sleeve 222 in response to the incline angle of theroller 23, and at that time, the incline angle of the roller 23 and thetravel of the spool 223 are determined by the pressure difference ΔP,not by the initial travel of the sleeve 222, so that, when the travel ofthe sleeve 222 is set so as for the communication relation between saidports to be kept even after the roller 23 is inclined and the spool 223is moved, or when the sleeve 222 is moved in the predetermined directionafter an initial travel thereof so as for the communication relationbetween said ports to be kept, the direct shift control by the pressuredifference ΔP is enabled even after the roller 23 is inclined and thespool 223 is moved.

In this transmission 10, the direct control is always implemented nearthe geared neutral condition where the influence of the hysteresis islikely to appear in the three-layers valve control, in other words, isimplemented during low vehicle speed. In addition, the control unit 300of the transmission 10 is made to implement the control which dare notmake the geared neutral condition in order to generate a creep force(hereafter, referred to as “creep control”) as an automatic transmissionhaving a torque converter when the vehicle speed is in the low speedrange, where the direct control is implemented, and the idle switch 308is on. The reason thereof will be described below.

The geared neutral means to keep the internal gear 53 of the planetarygear mechanism 50 stationary by setting the ratio between the rotationspeed inputted into the sun gear 52 of the planetary gear mechanism 50through the high mode gear train 90 and that inputted into the pinioncarrier 51 of the planetary gear mechanism 50 through the low mode geartrain 80 to the predetermined value, and therefor the toroidal speedratio is controlled by the three-layers valve control or the directcontrol described above, and there is only one rotation speed ratiobetween the sun gear 52 and the pinion carrier 51 to actualize thegeared neutral and therefore there is only one toroidal speed ratio. Asa result, extremely minute toroidal speed ratio control is required andit is shifted frequently to the forward or the backward runningdirection.

When the vehicle starts to move from the temporary stop condition, thegeared neutral does not allow the vehicle to start only be releasing thebrake pedal but requires to depress the accelerator pedal. Accordingly,to secure a good startability by always applying a certain degree ofdriving force to the vehicle as the automatic transmission with a torqueconverter does, the toroidal speed ratio must be controlled with someoffset from the geared neutral position, for example, so as to slightlyapply a forward driving force in the forward running range such asD-range, and so as to slightly apply a backward driving force in thebackward running range of the R-range. This kind of creep control doesnot require such a minute control, so that this brings some advantage inbraking actuation. As described above, in this transmission 10, sincethe creep control is implemented when the vehicle speed is in low speedrange, where the direct control is implemented, and the idle switch 308is on, the three-layers valve control is switched into the directcontrol and at the same time into the creep control when, for example,the vehicle speed is lowered while the driver releases the acceleratorpedal, and, on the contrary, when the vehicle speed is lowered with theaccelerator pedal depressed on the up-hill etc., normal shift control isimplemented based on the shift map under the direct control and then thecreep control begins when the accelerator pedal is released fordepressing the brake pedal.

While the vehicle is at a stop, the creep force is made as small aspossible to save the fuel consumption, and when starting, the creepcontrol is applied from the beginning, and then is replaced by thenormal direct control as the accelerator pedal is depressed, and, whenthe vehicle speed exceeds a certain level, the three-layers valvecontrol is applied.

(2) Concrete actuation in respective controls

As shown in FIG. 20, various control programs are stored in the controlunit 300 to cope with various kinds of conditions based on the shiftactuation described above, and the interruption by each control isexecuted when required independently or associated with other controls.

(2-1) Line pressure control

As described above, the pressure of the operating fluid discharged fromthe oil pump 102 is supplied to the main line 201 through the regulatorvalve 202 after being regulated by the linear solenoid valve 209 intopredetermined line pressure, but, in the shift control, this linepressure is led to the three-layers valve 220, 230 together with therelief pressure which is supplied to the relief pressure line 203through the relief valve 204 after being regulated by the linearsolenoid valve 210 into a pressure lower than the line pressure, and isused as an important pressure to generate the pressure difference ΔP forthe shift control in which, while the roller 23, 33 or the trunnion 25,35 of the continuously variable transmission mechanism 20, 30 being heldin the neutral position against the traction force T, the trunnion 25,35 is moved in the predetermined direction to incline the roller 23, 33.

Accordingly, the pressure difference ΔP is controlled so as to hold thetrunnion 25, 35 in the neutral position in response to the increase ordecrease of the traction force T, and, for example, when the reliefpressure is constant, the pressure difference ΔP can be expanded byincreasing the line pressure to counteract the larger traction force T,and, when the line pressure is constant, the pressure difference ΔP canbe expanded by decreasing the relief pressure to counteract the largertraction force T.

The traction force T is varied not only by the engine torque but also bythe incline angle of the roller 23, 33. As shown in FIG. 21 by anexample of the first roller 231 of the first continuously variabletransmission mechanism 20, when the roller 231 is inclined into thedeceleration side as a result of the shift control, as shown in thedrawing by a solid line, a radius r1 of the contact point between theroller 231 and the input disk 21 becomes smaller comparing with the casewhere the roller 231 is inclined into the acceleration side as shown bya chain line in the drawing, and, therefore, even if the torque Tztransmitted from the input disk 21 to the roller 231 is constant, thedrag force applied to the roller 231 at the contact point becomes largerand the reaction force at the contact point of the roller 231 with theoutput disk 22 also becomes larger. Thus, as the roller 231 inclinesinto the deceleration side, the traction force T increases as a whole.

The torque Tz is transmitted in the direction described above in thehigh mode (H-mode) where the low mode clutch is disengaged and the highmode clutch is engaged, and, in this high mode, as the speed ratio ofthe continuously variable transmission mechanism 20, 30 (hereafter, alsoreferred to as “toroidal speed ratio”) becomes larger, the line pressureis controlled to be increased when the relief pressure is constant, orthe relief pressure is controlled to be decreased when the line pressureis constant so as for the pressure difference ΔP for counteracting thetraction force T to be expanded.

On the other hand, in the low mode (L-mode), the torque is transmittedin the opposite direction of that of the high mode due to thecirculating torque circulated back to the continuously variabletransmission mechanism 20, 30 as a reaction force from the planetarygear mechanism 50 (see FIG. 9). Accordingly, in the low mode, when theroller 231 is inclined into the acceleration side as shown by a chainline in FIG. 21, a radius r2 of the contact point between the roller 231and the output disk 22 becomes smaller and thereby the traction force Tbecomes larger, and, therefore, as the toroidal speed ratio becomessmaller, the line pressure is controlled to be increased when the reliefpressure is constant, or the relief pressure is controlled to bedecreased when the line pressure is constant so as for the pressuredifference ΔP for counteracting the traction force T to be expanded.

The concrete actuation of the line pressure control by the control unit300 is shown, for example, in FIG. 22, wherein the engine torque Te iscalculated from the engine speed Ne and the throttle angle θ in stepS11, the oil pump loss “Loss” is calculated in step S12, and thetoroidal speed ratio Rtd is calculated from the input and the outputrotation speeds of the continuously variable transmission mechanisms 20,30 in step S13 respectively, and then the transmission torque Tz valueis determined in step S14 from the map shown, for example, in FIG. 23 bythe use of above calculated values and modes as parameters. As shown inthe map, in the low mode D-range, the transmission torque Tz increasesas the toroidal speed ratio Rtd moves to the acceleration side, and thetransmission torque Tz is fixed to 1.0 in the high mode.

Then, in step S15, the line pressure PL is determined based on thetransmission torque Tz from the map shown, for example, in FIG. 24, and,in step S16, the linear solenoid valve 209 is controlled so as for theline pressure PL to be obtained. In this map, the line pressure israised up in the range where the transmission torque Tz exceed apredetermined value in order to counteract the traction force T,wherein, as described above, the line pressure is set to become largeras the toroidal speed ratio Rtd moves to the acceleration side in thelow mode, and the line pressure is set to become larger as the toroidalspeed ratio Rtd moves to the deceleration side in the high mode. Theline pressure is fixed to a constant value in the range where thetransmission torque Tz is less than the predetermined value, and, inthis range, the pressure difference ΔP is controlled by increasing ordecreasing the relief pressure. That is, in the low mode, the reliefpressure is decreased as the toroidal speed ratio Rtd moves to theacceleration side, and, in the high mode, the relief pressure isdecreased as the toroidal speed ratio Rtd moves to the decelerationside.

(2-2) Engage control

As described above, since, in the N-range, the main line 201 forsupplying the line pressure is intercepted from the first to the thirdoutput lines 205-207 by the manual valve 208, both of the low mode andthe high mode clutches 60, 70 are in the disengaged condition. When thedriver shifts the range from this condition to the forward running rangesuch as D-range, S-range and L-range, or to the backward running rangeof R-range, the low mode clutch 60 is engaged to make the low mode. Atthat time, if the toroidal speed ratio is controlled to make that of thegeared neutral, the pinion carrier 51 of the planetary gear mechanism 50and the secondary hear 82 of the low mode gear train 80 are synchronizedin rotations with each other, so that if the low mode clutch 60 forconnecting or disconnecting them is engaged, the engagement shockscarcely occurs.

However, since the N-range is generally selected at a stop with idlingcondition or in low vehicle speed, the engage actuation of N-D range orN-R range is implemented during the creep control. Thus, since thegeared neutral condition is not employed in the creep control, theengagement shock is generated by the creep torque when the low modeclutch 60 is engaged.

The control unit 300 implements the engage control to suppress theengagement shock according to the flow chart shown in FIG. 26. Then theengage control will be described with reference to FIG. 26 which showthe relation between the pulse number of the step motor 251 and thefinal speed ratio, FIG. 27 which shows the relation between the reliefpressure and the output torque, and FIG. 31 which shows a time chart.

The control unit 300 judges, at first in step S21, if the range is inthe N-range or not in the previous control cycle, and, in case of YES,judge if the current range is a running range such as D-, S-, R-rangesin step S22. In case of NO, which means N-range is continued, a reliefpressure Prf is made to zero in step S23, and, in step S24, a pulse ofthe step motor 251 PLUS is set to PN which makes geared neutralcondition and then a timer value TIME is set to zero in step S25.

The reason why the relief pressure Prf is set to zero when the N-rangeis continued is that it provide an advantage that unnecessary power isnot consumed because the relief pressure Prf becomes zero when thelinear solenoid valve 210 for relief pressure control is not actuated.The reason why the pulse PLUSE is set to PN which makes geared neutralcondition is to make the sleeve 222 return to a set point as apreparation for generating a creep force by the direct control in theengage actuation expected thereafter, and therefore other point may beemployed if it makes a relation between the sleeve 222 and the spool 223of the three-layers valve 220 be in a predetermined neutral position andthereby communication between each port is intercepted.

On the other hand, when the current range is judged to be the runningrange such as D-, S-, L-, and R-ranges in step S22, the timer value TIMis judged in step S26, and when the timer value TIM is within apredetermined time TIMx which is required for engaging the low modeclutch 60, the relief pressure Prf is set to relatively highpredetermined pressure Prf(on) in step 27, and in order to keep thecommunicating condition between each port in the three-layers value 220to implement the direct control, in step S28, when the current shiftedrange is the forward running range such as D-range, the pulse PLUSE ofthe step motor 251 is shifted from PN to the PD1, where the final speedratio is in the high speed side, and, when the current shifted range isthe backward running range of R-range, the pulse PLUSE of the step motor251 is shifted to the PR1, where the final speed ratio is in the highspeed side, and then the timer value TIM is added by 1 in step S29.

That is, during the predetermined time TIMx required for engaging thelow mode clutch 60, the relief pressure Prf is set to higher value, andthereby the pressure difference ΔP, the offset from the line pressure,is made smaller to close to the geared neutral condition, and the creepforce (output force) is set lower. Therefore, the engagement shock inthe engage actuation is suppressed.

When the timer value TIM exceeds the predetermined time TIMx requiredfor engaging the low mode clutch 60 in step S26, the relief pressure Prfis set to relatively low predetermined pressure Prf(off) in step S31,and in order to keep the communicating condition between each port inthe three-layers valve 220 to implement the direct control, in step S32,when the current shifted range is the forward running range such asD-range, the pulse PLUSE of the step motor 251 is shifted from PN to thePD0, where the final speed ratio is in the low speed side, and, when thecurrent shifted range is the backward running range of R-range, thepulse PLUSE of the step motor 251 is also shifted to the PR0, where thefinal speed ratio is in the low speed side, and then the timer value TIMis set to zero in step S33.

That is, after the low mode clutch 60 is engaged, the relief pressure isset to lower value, and thereby the pressure difference ΔP, thedifference from the line pressure, is made large to expand the offsetfrom the geared neutral condition, and the creep force (output force) isset higher, Therefore, excellent startability may be secured.

(2-3) Direct control

Adding to the basic actuation of the direct control itself describedabove, the control unit 300 of this transmission 10 implements somespecial control when the brake pedal is depressed or during creepingspeed. The concrete control actuation in these cases is shown by theflow chart in FIG. 28, and this will be explained with reference to thetime chart in FIG. 31. At first in step S41, the three-layers valuecontrol is replace be the direct control when the vehicle speed V fallsbelow a certain speed which is higher than the target vehicle speed Voin the creep control by a predetermined value ΔV, and, at that time,when the brake switch 309 is judged to be ON in step S43 (at that time,the idle switch 308 is in on and the creep control is begun), the reliefpressure Prf is set to relatively higher predetermined pressure Prf(on)in step S43, and, in step S44, the linear solenoid valve 210 iscontrolled so as for this relief pressure Prf(on) to be obtained. Thatis, the relief pressure Prf is raised up to make the creep force smallersince an earlier deceleration is preferable when the brake pedal isdepressed.

On the other hand, when the brake switch 309 is in OFF in step S42, therelief pressure Prf is set to the relatively lower predeterminedpressure Prf(off) in step S45. When the idle switch 308 is in ON in stepS46, a deviation dV of the current vehicle speed V from the targetvehicle speed Vo in the creep control is determined in step 47, and thena feedback pressure ΔPrf of the relief pressure Prf is derived from themap in FIG. 29 based on the deviation dV in step S48. The reliefpressure Prf added by the feedback pressure ΔPrf is determined in stepS49, and, in step S44, the linear solenoid valve 210 is controlled so asfor this relief pressure Prf to be obtained. Thereby, the creep force isnot decreased when the brake pedal is not depressed, so that the vehiclespeed can be kept in the target vehicle speed Vo by the feedbackcontrol.

The time charge of FIG. 31 shows the feedback control of the vehiclespeed to the target vehicle speed Vo in a stop and during starting. Thereason why the initiating condition of the direct control is set to thecertain speed, which is higher than the target vehicle speed Vo by apredetermined value ΔV in step S41, is to prevent the direct controlfrom being replaced by the three-layers valve control due to theovershoot during the feedback control of the vehicle speed V.

When the idle switch 308 is in OFF, which means that the acceleratorpedal is depressed, in step S46, the relief Prf is determined inresponse to the throttle angle θ in step S50, and, in step S44, thelinear solenoid valve 210 is controlled so as for the relief pressurePrf to be obtained (during Δt period in the vehicle starting in FIG.31). At that time, the relation between the relief pressure Prf and thethrottle angle θ is set in a map, as shown in FIG. 30, so that therelief pressure Prf increases as the throttle angle θ increases.Thereby, the more the accelerator pedal is depressed, the more the creepforce is lowered, that is, it moves close to the geared neutralcondition and, as a result, the speed ratio becomes larger, the enginespeed is raised, a better acceleration performance is achieved, and theswitching to the three-layers valve control is performed smoothly.

When the vehicle speed V exceeds the speed of initiating condition ofthe direct control in step S41, the relief pressure Prf which generatesthe pressure difference ΔP associated with the line pressure in thethree-layers valve control is set to zero in step 51, and, in step S52,the linear solenoid valve 210 is controlled so as for this reliefpressure Prf is obtained and, in step S53, the control system moves tothe three-layers valve control.

The pulse number of the step motor 251 in the point of switching betweenthe three-layers valve control and the direct control does notnecessarily coincide with each other, and, when the direct controlstarts, the sleeve 222 is moved from the position where the three-layersvalve control ends to the corresponding position of the direct control(pulse number PD0), and, when the three-layers valve control starts, thesleeve 222 is moved from the position where the direct control ends(pulse number PD0) to the corresponding position of the three-layersvalve control.

In the direct control, when the brake switch 309 is in ON in step S42,the relief pressure Prf is increased to make the creep force smaller, sothat there exists some possibility of reverse direction running due tothe decrease of the driving force when the creep force is immediatelylowered upon switching ON of the brake switch 309 in the case where thevehicle is stopped on the slope, not on the flat road. To cope with thisproblem, a second direct control program is stored in the control unit300 of this transmission 10.

The second direct control including this incline angle control will bedescribed with reference to the flow chart in FIG. 32 and the time chartin FIG. 35. The flow chart in FIG. 32 is the same with that in FIG. 28with an exception that step S40 is added before step S41 and the stepS43 is modified.

Prior to the judgement of the starting or the ending condition of thedirect control in step S41, a delay time Tc and the relief pressure Prfare determined in response to the road incline angle “k” detected by theincline angle sensor 310 in step S40. At that time, as shown in FIG. 33,as an upward incline angle becomes steeper, the delay time Tc is set tobe longer and the relief pressure Prf is set to be lower (so that thecreep force becomes larger). The relief pressure Prf0 on the flat roadis set to the value which makes normal creep force.

The direct control starts at step S41, and, when the brake switch 309 isin ON in step S42, a count number “count” is judged whether it is zeroor not in step S43 a, and when the judgement is YES, that is, when it'sa first approach to the step S43 a, the relief pressure Prf (, which isdetermined based on the incline angle,) is set to the relatively lowerpredetermined pressure Prf(off) in step S43 b, as is done when the brakeswitch is in OFF, and then the count number is added by 1 in step S43 c,and the count number is compared with the delay time Tcd determinedbased on the incline angle in step S43 d.

When the count number is within the delay time Tcd, the above relativelylower predetermined pressure Prf(off) is kept in step S43 e, and, whenit exceeds delay time Tcd, the a calculation for increasing the reliefpressure Prf in response to the count number is executed in step S43 f.A correction coefficient Ck used in the calculation is set, as shown inFIG. 34, so that the steeper the incline angle is, the smaller the Ckis, that is, the more slowly the relief pressure Prf increases (namely,the more slowly the creep force decreases). The linear solenoid valve210 is controlled so as to obtain the relief pressure Prf determined asdescribed above.

According to this control, when the upward incline angle of the roadsurface is steeper, the creep force after the brake pedal beingdepressed is set to be larger and the delay time, which is a holdingtime thereof, is set to be longer, and, after the delay time havingpassed, the creep force is decreased more slowly when the upward inclineangle is steeper, so that the vehicle can be prevented effectively fromrunning into the reverse direction on the road with an incline angle.

(2-4) D-R switching control

During garage parking operation, for example, the shift range may bechanged from D-range to R-range (D-R) for expected backward runningwhile the vehicle still running forward, and also may be changed inopposite direction (R-D) while the vehicle still running backward. Asregards the gear train of this transmission 10 at that time, though themanual valve 208 passes through the N-range position while movingbetween D-range and the R-range, the low mode clutch 60 is kept in theengagement condition because of its short time.

The toroidal speed ration changes under this condition crossing thegeared neutral, that is, at that time, the toroidal speed ratio iscontrolled to change the rotation speed of the sun gear 52 so that theinternal gear 53 and the secondary shaft 13 are rotated in the oppositedirection to that of the current one. Since a large force is required toincline the roller 23, 33 of the continuously variable transmissionmechanism 20, 30 with respect to the disk 21, 22, 31, 32 in such manneras described above, there exists, as a result, a possibility of makingslip and damage on the roller 23, 33, disk 21, 22, 31, 32 and the like.

To cope with this problem, the control unit 300 implements a controlaccording to the flow chart shown in FIG. 36 in order not to apply highload onto the continuously variable transmission mechanism 20, 30 duringthe switching control between the forward and the backward running.

When the range is judged to be in D-range in step S61, a normalthree-layers valve control is applied by the sleeve movement based onthe shift diagram (shift map), for example, shown in FIG. 18 in stepS62, and, when the range is not in D-range in step S61 and is in N-rangein step S63, the low mode clutch 60 is disengaged in step 64, and thesleeves 222 of the three-layers valve 220 is moved to the position closeto the geared neutral in step S65, and then a original positioncorrection of the step motor 251 is implemented in step S66. The reasonwhy the sleeve 222 is not moved to the geared neutral position but tothe position close to the geared neutral in step S65 is that it isdifficult to move the sleeve 222 to the exact position of the gearedneutral as described above, and it is needless to say that the sleevemay e moved to the geared neutral position (the position to which thesleeve 222 is moved in step 65 will be referred to as “referenceposition”).

Thereby, in the N-range, the power transmission path is intercepted andthe sleeve 222 is moved to the reference position, and, at that point,the original position correction of the step motor 251 is implementedagain. The original position correction of the step motor 251 isimplemented in the following procedure. At first, the toroidal speedratio in the case where the sleeve 222 is in the reference position iscalculated based on the detected values of the input rotation speedsensor 306 installed on the low mode clutch drum 61 and the outputrotation speed sensor 307 installed on the second gear 92 of the highmode gear train 90. The pulse number when the sleeve 222 is moved to thereference position is set to the original position pulse number (forexample, it is near 1360 in FIG. 17). The calculated actual speed rationof the toroidal is compared with the predetermined ideal speed ratio ofthe toroidal in the reference position, and the sleeve 222 is moved tothe direction where the offset therebetween will be removed. Themovement of the sleeve 222 is implemented under feed-forward control,and, after the sleeve 222 is moved by several pulses, the originalposition pulse number is replaced by this pulse number of the step motor251.

Referring to FIG. 36 again, when the range is not in the D-range in stepS61 and is not in N-range in step S63, it is judged whether it is in theR-range or not in step S67, and when it is NO, since it must be eitherof S-range or L-range, the step is progressed to step S62, and, when itis YES, it is judged whether the vehicle is running backward or not inStep 68. When the vehicle is running backward, the normal three-layersvalve control is applied in step 62, and when it is NO, it is judgedwhether the vehicle speed is zero or not in step 69, and when YES, whichmeans that the vehicle is running forward in a certain speed, each stepof steps S64-66 which are implemented in the N-range is implemented. Onthe contrary, when it is NO in step S69, which means that the vehicle isat a stop with the R-range, the step is progressed to step 70, and thesleeve 222 of the three-layers valve 220 is moved to the reverse startposition. More concretely, it is moved to a creep start position wherethe internal gear 53 and the secondary shaft 13 is rotated into thebackward running direction. Then the low mode clutch 60 is engaged instep S71.

According to this control, when the switching to the R-range is operatedduring forward running condition, the step is progressed through thepath of S61, S63, S67, S68, and S69, and, after the low mode clutch 60is disengaged in step S64, the vehicle stop is confirmed in step S69,and then the sleeve is moved to the backward running direction in step70, and the low mode clutch 60 is engaged in step S71, and thereby, thesun gear 52 of the planetary gear mechanism 50 rotates under light loadwhile the low mode clutch 60 is disengaged, and, during this period, theroller 23 of the continuously variable transmission mechanism 20 is madeto be inclined so as to change the rotation speed of the sun gear 52, sothat the inclining motion can be implemented under low load, and therebythere exist no fear of causing the slip and damage on the roller 23, 33and the disk 21, 22, 31, 32.

(2-5) R-D switching control

Though the flow chart shown in FIG. 36 represents the control of the D-Rswitching, the same method can be applied to the R-D control. Thecontrol flow thereof is shown in FIG. 37.

(2-6) Shift control during backward running

The continuously variable transmission 10 can control the toroidal speedratio continuously, and, thereby, the final speed ratio can bearbitrarily changed from the geared neutral to either of forward orbackward direction by changing the rotation speed of the sun gear 52.Therefore, though an infinite number of gear positions can be set forthe backward running, a considerable caution is required while startingespecially in the backward running, which is different from the case ofthe forward running which requires good accelerating ability fromstandstill.

Then, the control unit 300 of this continuously variable transmission10, as shown in FIG. 38, makes a shift control, when the range is inR-range in step S101, by the use of the shift map for backward runningin step S102, and, when the range is in D-range in step S101, by the useof the shift map for forward running in step S103.

At that time, as shown in FIG. 39, the shift map for the backwardrunning determines lower engine speed value as a target value Neocomparing with that for the forward running for the same vehicle speed Vand the same throttle angle θ. In other words, the final speed ratio isshifted to the high speed side as a whole, and thereby too quickstarting can be avoided in the backward running.

This characteristic of the shift for backward running may be appliedonly to the speed range lower than the predetermined one. At that time,the backward running in the same final speed ratio with that of theforward running is enable except the starting where a special caution isrequired.

FIG. 39 does not show the shift characteristic below a vehicle speedVo+ΔV, which is used in judgment in the direct control, because thischart is made based on the time chart shown in FIG. 31 where the idleswitch 308 is already in ON when the three-layers valve control isswitched to the direct control, and thereby the creep control startsimmediately, so that the normal shift control is not employed below thevehicle speed for judgement Vo+ΔV, that is, this kind of shift map isnot used.

(2-7) Low mode/high mode switching control

As described with reference to FIG. 17, the low mode characteristiccurve and the high node characteristic curve of the D-range cross eachother in the predetermined pulse number of the toroidal speed ratio.This is represented by a mode switching line in the shift map shown inFIG. 18 or FIG. 39. That is, the low mode clutch 60 and the high modeclutch 70 are switched at a point where the final speed ratio of bothmodes coincide each other. Thereby, both modes can be switched eachother without making shock due to sudden change in the final speedratio.

Since the switching between both clutches 60, 70, however, takes acertain period of time, the vehicle running condition might not be onthe mode switching line when the switching is completed, and, as aresult, it may make a sudden change in speed ratio.

To cope with this problem, the control unit 300 implements a modeswitching control according to the flow chart shown in FIG. 40. At firstin step S111, the control unit 300 judges whether or not the actualengine speed Ne detected by the engine speed sensor 302 is on theapproach to the value determined by multiplying the final speed ratio Goof the mode switching line by the vehicle speed V detected by thevehicle speed sensor 302. That is, it judges whether the current finalspeed ratio is nearly equal to that on the mode switching line or not.

When it is YES, in step S112, the toroidal speed ratio is controlled soas for the current final speed ratio G to be kept while the clutches 60,70 being switched. Then in step S113, a deviation ΔN of the actualengine speed Ne from the target engine speed Neo required to keep thecurrent final speed ratio G is calculated, and, in step 114, a feedbackvalue ΔPLUS of the pulse PULSE is determined from the map shown in FIG.41 which is set so as for the engine speed deviation ΔN to be made zero,and finally in step S115, the feedback value ΔPLUS is outputted to thestep motor 251.

Thereby, the position of the sleeve 222 of the three-layers valve 220 isfeedback-controlled, the engine speed deviation ΔN is made to be zero,and, as a result, the final speed ratio is fixed to a constant value.Since the modes are switched during this period, there is no change inspeed ratio before and after switching modes, so that, the modes can beswitched smoothly without shock.

According to the present invention, since, in the toroidal typecontinuously variable transmission for the front engine front wheeldrive vehicle, which employs geared neutral system, the gear train whichtransmits the rotation to the second shaft side on which the planetarygear mechanism is mounted is disposed on the opposite side end, withrespect to the engine, and the first shaft on which the continuouslyvariable transmission mechanism is mounted, this gear train can beprevented from interfering with the differential gear unit, which isengaged with the engine side ends of the second shaft, or the powertransmission mechanism to said unit. Therefore, the length in the axialdirection of the transmission can be made shorter comparing with thecase where the gear train is disposed with an offset from thedifferential gear unit and the like in the axial direction, so that themounting operation to a vehicle body as well as a layout design of thetransmission can be improved.

According especially to the fourth invention of the present invention,since, in the construction where two continuously variable transmissionmechanism are disposed on the first shaft, the loading mechanisminterposed between the first shaft and the input portion of thesecontinuously variable mechanisms is disposed on the opposite side end,with respect to the engine, of the first shaft as the above gear train,the circulating torque which is generated by the planetary gearmechanism on the second shaft and is circulated back to the first shaftside under the geared neutral or the low mode condition is nottransmitted to the first shaft, so that the first shaft is required nomore than to have a diameter or the strength for transmitting the torquefrom the engine, and, as a result, the transmission is allowed to bemade with lower cost, to be compact, and to have an improved durabilityand a lighter weight, and, at the same time, the vibration and the noiseon the vehicle can be reduced since the vibration from the engine can beeffectively absorbed due to the reduced rigidity of the first shaft.

What is claimed is:
 1. A toroidal continuously variable transmissioncomprising: a first shaft whose one end is coupled with an engine; asecond shaft which is disposed in parallel with said first shaft andwhose engine side end is coupled with a differential gear unit fordriving a left and a right driven wheels; a toroidal continuouslyvariable transmission mechanism, disposed on said first shaft, includingan input disk coupled with said first shaft, an output disk which isdisposed on an engine side of said input disk and is rotatably supportedon the first shaft, a roller interposed between said both disks fortransmitting power therebetween, and a contact point control memberwhich inclinable supports said roller and varies a speed ratio betweenboth disks by changing contact points between said roller and said inputand said output disk; a planetary gear mechanism, disposed on saidsecond shaft, including three rotary elements of a sun gear, an internalgear and a pinion carrier in which among these rotary elements, thefirst element is coupled with said output disk of said continuouslyvariable transmission mechanism so as to rotate together therewith andthe second element is coupled with the second shaft; a gear trainincluding a first gear which is disposed on the first shaft at anopposite side, with respect to the engine, of the continuously variabletransmission mechanism for a rotatable movement together with the firstshaft, a second gear which is rotatably supported on the second shaft atan opposite side, with respect to the engine, of the planetary gearmechanism, and an idle gear which is engaged with said first and saidsecond gears to transmit a power therebetween; a first clutch mechanismfor engaging or disengaging said second gear of said gear train with thethird element of the planetary gear mechanism; a power transmission pathfor transmitting a driving torque from said output disk to the secondshaft without passing it through said planetary gear mechanism; a secondclutch mechanism for engaging or disengaging said power transmissionpath; and a control unit for controlling an operation of said contactpoint control member, said first clutch mechanism, and said secondclutch mechanism.
 2. A toroidal continuously variable transmission asrecited in claim 1 further comprising a vehicle speed sensor, whereinsaid control unit controls said first clutch mechanism so as to engagesaid second gear with said third element and controls said second clutchmechanism so as to disengage said power transmission path when a vehiclespeed is lower than a predetermined vehicle speed, and, on the otherhand, controls said first clutch mechanism so as to disengage saidsecond gear from said third element and controls said second clutchmechanism so as to engage said power transmission path when the vehiclespeed is higher than the predetermined vehicle speed.
 3. A toroidalcontinuously variable transmission as recited in claim 2 furthercomprising an engine load detector, wherein said predetermined vehiclespeed is set to be higher as an engine load is increased.
 4. A toroidalcontinuously variable transmission as recited in claim 1, wherein saidtoroidal continuously variable transmission mechanism further comprises,in addition to a first continuously variable transmission mechanismcomprising the input disk coupled with the first shaft, the output diskwhich is disposed on an engine side of said input disk and is rotatablesupported on the first shaft, the roller interposed between said bothdisks, and the contact point control member for changing contact pointsbetween said roller and said input or said output disk, a secondcontinuously variable transmission mechanism comprising a second outputdisk which is disposed on an engine side of the output disk of saidfirst continuously variable transmission mechanism and is rotatablysupported on the first shaft, a second input disk which is disposed onthe engine side of said second output disk and is coupled with the firstshaft, a second roller interposed between said both second disks, and asecond contact point control member for changing contact points betweensaid second roller and said input and said output disk, wherein theoutput disk of said first continuously transmission mechanism and theoutput disk of said second continuously transmission mechanism areintegrated into one unit and a gear is formed on an outer surface ofsaid integrated output disk unit for engaging and rotating saidintegrated output disk together with the first element of the planetarygear mechanism with each other.
 5. A toroidal continuously variabletransmission as recited in claim 4, wherein said toroidal continuouslyvariable transmission mechanism further comprises, in addition to saidfirst continuously variable transmission mechanism comprising the inputdisk coupled with the first shaft, the output disk which is disposed onan engine side of said input disk and is rotatably supported on thefirst shaft; the roller interposed between said both disks, and thecontact point control member for changing contact points between saidroller and said input or said output disk, the second continuouslyvariable transmission mechanism comprising the second output disk whichis disposed on the engine side of the output disk of said firstcontinuously variable transmission mechanism and is rotatably supportedon the first shaft integrally with said second output disk, the secondinput disk which is disposed on the engine side of said second outputdisk and is coupled with the first shaft, the second roller interposedbetween said both second disks, and the second contact point controlmember for changing contact points between said second roller and saidinput and said output disk, wherein said first shaft is inserted into athird shaft having a through-hole therein, and each of the input and theoutput disks of said first and said second continuously variabletransmission mechanisms is disposed on said third shaft, one end of saidthird shaft being supported by a transmission case through a bearing,the other end of said third shaft being fitted into the first gear ofthe gear train, said first gear being supported by a transmission casethrough a bearing, and a spring member is interposed in a fittingportion of said third shaft and the first gear for absorbing an axialrelative displacement therebetween.
 6. A toroidal continuously variabletransmission as recited in claim 4, wherein said toroidal continuouslyvariable transmission mechanism further comprises, in addition to saidcontinuously variable transmission mechanism comprising the input diskcoupled with the first shaft, the output disk which is disposed on anengine side of said input disk and is rotatably supported on the firstshaft, the roller interposed between said both disks, and the contactpoint control member for changing contact points between said roller andsaid input or said output disk, the second continuously variabletransmission comprising the second output disk which is disposed on theengine side of the output disk of said first continuously variabletransmission mechanism and is rotatably supported on the first shaft,the second input disk which is disposed on the engine side of saidsecond output disk and is coupled with the first shaft, the secondroller interposed between said both second disks, and the second contactpoint control member for changing contact points between said secondroller and said input and said output disk, wherein said first shaft isinserted into a third shaft having a through-hole therein, and theoutput disks of the first and the second continuously variabletransmission mechanism are integrally and rotatably supported on amiddle of said third shaft, and the input disks of the second and thefirst continuously variable transmission mechanisms are respectivelydisposed on the engine side and the opposite side, with respect to theengine, of said output disks and are coupled with said third shaft, anda loading mechanism for pressing the rollers by and between the inputand the output disks in the first and the second continuously variabletransmission mechanisms is disposed between the input disk of the firstcontinuously variable transmission mechanism and the first gear of thegear train disposed on the opposite side thereof with respect to theengine.
 7. A toroidal continuously variable transmission as recited inclaim 6, wherein said loading mechanism comprises a pair of disks whosesurfaces facing with each other are formed into cam surfaces havingcircumferential concave and convex shapes, and a roller which isinterposed between both disks to generate axial force between them by arelative rotation therebetween, and a pin member is interposed betweenthe first gear of the gear train and the disk located in said first gearside to integrally rotate them, said pin member being disposed in aportion where a thickness of the disk located in said first gear side israther thicker due to the concave and convex figures thereof.
 8. Atoroidal continuously variable transmission as recited in claim 1,wherein two oil channels for supplying the first clutch mechanism andthe second clutch mechanism with a coupling fluid respectively areprovided in the second shaft, and said both oil channels are led from aside portion where a hydraulic pressure source is provided.